If you haven’t noticed, the website has a new domain name (web address).
This is a registered address that belongs to me. Having a registered address is more professional and it makes the website easier to find in search engines. It is just another ongoing development in the website. I created this website and constantly update and develop it myself. I am certainly not a professional website builder, but if you are someone wanting to build your own website, I would be glad to offer any help to you.
Be sure to update your bookmark to this new address. The system will automatically forward you from the old address until you change.
Check back, I will have a new technical post soon.
It seems like each person has their own way of adjusting valve lash. As with hydraulic lifters, talked about in the previous post, there is nothing mysterious or complicated about adjusting the valve lash. The valve being adjusted must be fully closed. That is the only rule that matters. Whatever procedure you want to use is up to you. The content of this post will be about the lash itself and not the adjusting procedure.
Since the solid lifter does not have the ability to take-up the clearance created in the valve train, there will always be some running clearance in a hot engine. Most engines will gain clearance in the valve train from a cold to a hot engine. I’m sure there are exceptions, but I have not come across one. This gain in clearance is taken into consideration when designing the cam profile ramps for a particular engine combination. A cast iron push rod engine will not gain as much clearance as an all aluminum push rod engine. The ramps should be designed differently for each application. Many times a cam profile is used without any thought given to what the actual running lash was designed to be. The more lash, the less area the cam profile has. The opening and closing valve velocity points will also increase which could cause mechanical damage to the valve train. Less lash will increase the seat-to-seat duration, which may bleed-off cylinder pressure. Changing the rocker arm ratio also changes the lash setting. There is more to choosing the correct cam profile than just lift and duration numbers.
The valve lash settings come from the cam profile design. It is not just some number that is made up. The material of the engine block and heads, the rocker arm ratio that will be used, and the application, are all considered when designing the cam profile ramps. Older cam profiles were mostly designed for cast iron engines with 1.5 or 1.6 rocker ratios. Many engines today are all aluminum or have aluminum heads and higher rocker arm ratios. Again, the best performing cam profiles will be the ones that are designed specifically for your combination and application.
I will finish this post the same as the previous post. I know all kinds of combinations and lash adjustments are used for camshafts and lifters. I have already stated many times the disadvantages and dangers of mismatching cam profiles and lifters. I will just continue to shake my head and be amazed at why this is done.
Standard hydraulic lifters are designed to automatically compensate for any clearance that is created in the valve train from a cold engine to a hot operating engine. They will also compensate for any clearance that is created from wear or unintentional conditions. The lifter not following the cam profile is the most common unintentional condition. Lifter float or valve float is the more common term. This can be caused from not enough valve spring pressure, over revving the engine, or the wrong cam profile design. When this happens, the hydraulic lifter will “pump-up” to take up the temporary clearance created. Since the initial hydraulic lifter adjustment was set to operate at zero clearance, this extra “pump-up” of the lifter will keep the valve from fully closing until the lifter “bleeds-down” to its original setting. Until the lifter “bleeds-down”, that cylinder will obviously loose compression and power.
Special “anti-pump-up” hydraulic lifters are designed to keep the lifter from “pumping-up” during this condition. There is a strong snap ring retainer in the top of the lifter that stops the lifter plunger travel. The standard hydraulic lifter also has a retainer, but it is usually a weak round wire type. It is not designed to resist the plunger movement. The “anti-pump-up” lifters are adjusted differently than the standard lifters. Standard hydraulic lifters have around a 0.200 total plunger travel. I still do not understand why they are designed with that much travel. Certainly much more then will ever be necessary. Probably to compensate for the “stacking” tolerances (engine, head, pushrod, etc.) in production. The preload adjustment can be set anywhere in this travel range. Normally, the plunger preload adjustment is less than 0.050 when done by hand. All that is needed, is enough preload to compensate for any clearance that is created in the valve train from a cold engine to a hot operating engine and from any wear. In reality, on a cast iron stock engine, this would be less than 0.010. An all aluminum engine may require more. It always amazes me how some mystery is created when adjusting hydraulic lifters. Take a hydraulic lifter apart and analyze it. Look closely at the oil holes, oil channels, check valve, and envision how the oil flows. Make sure you understand how it works. Simple… right? Certainly, no mystery involved. I am referring to a stock OEM lifter here. Special “anti-pump-up” lifters operate the same way as stock ones. The plunger travel is only around 0.050 and they have that strong snap ring retainer. Adjustment is a little different. The idea is to preload the lifter so the plunger will just be touching the retaining ring during hot engine operation. This will prevent the lifter from “pumping-up” any further if valve float was to happen.
The retainer in the standard hydraulic lifter is not designed to resist the plunger travel, so the lifter should be preloaded enough to keep the plunger from touching the retainer during hot engine operation. The preload can be set anywhere within the plunger travel range. I guess this is where the mystery comes in. Usually you hear “quarter turn”, “half a turn” or “one full turn” on the rocker arm nut used to describe the preload adjustment. The plunger can also be adjusted where it is bottomed-out in the lifter. For engines without adjustable rocker arms, the pushrod length should preload the lifter at the midpoint of the plunger travel. All of these will work fine. Take your pick.
Cam profiles are designed a specific way to use a hydraulic lifter. I know all kinds of combinations and lash adjustments are used for camshafts and lifters. I have already stated in previous posts, the disadvantages and dangers of mismatching cam profiles and lifters. I will just continue to shake my head and be amazed at why this is done.
This is the new website for Ingram Engineering Cam Profile Designs. The old website service is being discontinued and that website will no longer be updated and will eventually be deleted. Be sure to change your bookmark to this new website address. The new service is much easier to use and has more features. Hopefully, everyone will like the results. Please leave a comment.
The majority of my cam profile designs are for conventional valve trains. Meaning the valve lift is a constant ratio compared to the cam profile lift. An example would be if the rocker arm ratio were 1.5 to 1, the valve lift would be the cam lobe lift multiplied by 1.5 at any point. Push rod engines with a rocker arm and overhead cam engines with a direct acting tappet are examples of a conventional valve train. Unconventional valve trains have a ratio that will vary through the valve movement or a cam lobe shape that is not typical. Overhead cam engines with finger followers are an example of a ratio that will vary. A Desmodromic valve train is also unconventional and very interesting. I would recommend studying this valve train. There are a few other older valve trains that are unconventional, but are not used in modern engines. They are still worth studying. You cannot fully understand something without knowing the history and its evolvement. Early internal combustion engines did not have an intake cam lobe and only used an exhaust lobe that opened at the bottom of the exhaust stroke and closed at the top. The intake valve used atmospheric pressure and a valve spring to open and close. To go from no intake cam to what we have today, is an interesting and educational journey.
Different cam profile design programs are written for the different valve train designs. Each program is costly and has to be justified based on the demand for certain profile designs. Much of the input data for these unconventional cam profile designs consists of angles and dimensions for that particular valve train. This data is usually not known and is not easily measured. Having the program only solves part of the problem. Usually a blueprint of that valve train is also necessary. Good luck getting that! Many times, if the customer can plot a ratio table of cam lift to valve lift in 1-degree increments, I can design a cam profile.
Like all of the posts, be sure to read the previous ones first.
I hope some of you participated and drew the lift curve in the previous post. For those that did, you will be further along in understanding cam profiles. If you haven’t already, play around with creating your own lift curves. There are many important observations that can be made by creating different lift curves.
An increase in the slope of the straight-line segment will increase the area of the profile. Once the lobe lift and the duration at 0.050 are chosen, the slope of this line will be the main difference between profiles with the same lift and 0.050 duration numbers. Would you intentionally design a cam profile with less area than what is possible? If you want less area, would it not be better to just design a cam profile that is smaller at 0.050 and/or less lift, but still with as much area as possible. Common sense will give you the answer. This is where you have to watch out for the marketing gimmicks.
Another observation is the length of the ramp area. The ramp area will be from 0.000 to 0.020. The valve contact point will fall somewhere in this range. This 0.020 can be divided-up into as many degrees as you would like, but after a certain length, it is just senseless to keep going. Around 20-degrees or less is plenty of room for any type of ramp design.
You will notice how the duration spread between 0.020 and 0.050 will set-up the slope for the lift curve. The terms high intensity and low intensity cam profiles are more of a marketing term, but relates to this spread. The closer the spread, the steeper the slope will be and more area will be created. This gives you more of a practical explanation without the marketing.
The top of the lift curve coming from maximum lift should be a nice, gentle curve. There should be no dwell at maximum lift and no corners blending into the straight-line segment of the lift curve. The lift curve from 0.050 to 0.020 to 0.000 should also be a nice, gentle curve. After you draw a few lift curves free handed, you will be able to see them in your mind and do it with your eyes closed. I know that cam profiles are sometimes designed with a dwell at maximum lift. These are special application profiles only and are not considered a good design.
Being able to look at a lift curve and seeing this information will be a big advantage when comparing cam profiles or analyzing a particular cam profile. It also takes something that seems very complicated and breaks it down into something simple and easy to understand.
Here is a simple exercise to start with. We are going to create a cam profile in the form of a lift curve chart. Draw a horizontal and a vertical line. These will be the “x” and “y” axis of the chart. The vertical line is the cam profile lift and the horizontal line is the degrees of rotation (duration). Zero is the point where both axis meet. This chart represents only the opening side of the cam profile. Next, plot the maximum lift of the profile. Now, plot the duration points at the following lifts: 0.020, 0.050, 0.100, 0.200, and 0.300. In the example, the maximum lift is 0.400. The plotted duration points are 68 (0.020), 60 (0.050), 52 (0.100), 39 (0.200), 26 (0.300). Remember, this is camshaft duration, not crankshaft. This is also only half of the cam profile, so the duration points are ¼ of the crankshaft duration. If this doesn’t make sense, think about it until it does. It will help to go back and study the posts on lift tables. From 0.020 to 0.000 will be the ramp area. The beginning of the ramp at 0.000 is 90-degrees. The chart should look like the one below. The plotted points show up as the red dots.
Now, with a smooth, gentle curve starting at 0.400, connect the plotted points down to 0.000 at 90-degrees. If you have been drawing along with your own chart, guess what? You have basically just created a cam profile and you didn’t even have to use any mathematical equations. Obviously, there is more to designing a cam profile then just this, but this is a legitimate profile with good area that would perform well if manufactured. If you haven't been playing along, you can take a short cut and print the chart.
Now, let’s analyze the lift chart closer. Notice the curve you drew connecting the plotted points is straight between 0.300 and 0.100. Depending on the profile lift, this straight segment will be shorter or longer in length. There will always be a straight segment in a good profile design. This straight segment has a slope to it, right? Maybe around 50-degrees with the horizontal axis. As the slope of this line increases, the area will also increase. The velocity and acceleration numbers will also increase. At some point, the slope of this line will be at maximum for the profile. When this maximum slope is reached, the cam profile is what I call “maxed out”. Basically, all of the area has been designed into the profile, within the limitations. It would be difficult to “better” this type of profile. Trying to make the profile smoother would be the only possible outcome.
I create a lift chart many times, as the first step in designing a cam profile. Especially, if a customer is giving me durations at different tappet heights. I will know right away, by the slope of the straight-line segment, if the cam profile will even be possible. I will usually plot duration points at 0.020, 0.050, and then at every 0.100 lift interval. From 0.020 down to 0.000 is the ramp area. The length of the ramp is usually around 20-degrees. In the example, the ramp length is actually 22-degrees. A good modern solid or hydraulic ramp can be designed with 20-degrees or less of length. You can see in the chart, a longer ramp will serve no purpose. It only keeps the valve separated from its seat for a longer period of time. No advantage. There is no significant airflow in this area. The area under the lift curve is that area inside the curve drawn to connect the plotted points. It is easy to see how changing the curve, changes the area. The area to increase for better performance is that above 0.050. Not below.
Usually, a cam profile concept starts with a maximum lift and the duration at 0.050. From there, the duration at 0.020 is around 30-degrees bigger than at 0.050. That spread determines the high and low intensity profiles that you hear talked about. You can see on the chart how the 0.020 and 0.050 durations will setup the slope for the straight-line segment.
Much can be learned from the cam profile lift chart. Whether from an existing profile or the design of a new one. The lift chart is essentially the cam profile.
It is difficult to find practical information on cam profiles. Mostly, all kinds of mathematical formulas and equations will show up when searching cam profile designs. The math doesn’t really tell you anything about the cam profiles. I try to give sincere, practical information about designing cam profiles in these blog entries. If you read and understand all of the entries so far, you will have a solid foundation of knowledge on the subject. Certainly more than the average engine person.
I cannot imagine anyone designing cam profiles today that does not use a computer. The computer does all the math that I mentioned earlier. A person still has to create the profile and determine if it is the proper profile for the application. That type of knowledge is difficult (if not impossible) to find. Mostly, it comes from many years of experience. It is not common knowledge and is not an exact science like the mathematical part is.
The marketing of cam profiles is usually directed toward the application side. Some marketing is occasionally directed at the mathematical techniques that are used to design the cam profile. As the average engine person becomes more knowledgeable and has access to more technical equipment, the less the marketing gimmicks work.
While I was at Reed Cams, I was able to profile every camshaft that I could. Before we had a computer profiler, I would use a manual one. We kept pages and pages of cam profiles for reference. Later, we kept files and files stored in the computer. Basically, all this data was knowledge to me. It showed me different design techniques and was a good way to compare profiles from different manufacturers. After you see a certain number of cam profiles, the marketing gimmicks are ignored.
I would say that anyone in the business of designing cam profiles today could create a good profile. It is rare to see a “bad” profile designed today using computers. Some profiles will certainly make more power than others and some profiles will create more stress on the valve train than others. The problem is that new cam profiles are not used. Older profiles, improper profiles, and copied profiles are still used way too often. See the last blog entry.
Keep reading my blog, ask questions, expand your knowledge, and don't let the marketing gimmicks be what guides you.
It’s 2015, but for the typical engine builder and racer, much of the same stuff is still going on that I remember happening twenty years ago or even longer. Hydraulic lifters are run on solid cam profiles, solid lifters are run on hydraulic cam profiles, valve lash is set with no regard to the cam profile design, camshaft grinders are still copying lobe profiles, base circle diameters are ground to any size with no regard to the original profile design, roller cam profiles are not designed for the roller wheel diameter used, and on and on. All of the stuff that ticked me off years ago is still going on. It’s sad.
With CNC cam grinding, machining, and head porting being common today, I would think these types of tactics with camshafts would be gone by now. Much of the blame should go to the racing engine rules that have to be dealt with. Hey…NASCAR is finally allowing roller camshafts in the Cup engines this year. WOW!
Performance cam profile designs are mostly defined by the engine design and the materials of the day. The end user’s perception, whether right or wrong, also plays a part. I have many cam profile design ideas (as do other cam profile designers) that are just not applicable to today’s racing engines or what the market wants.
In the OEM production world, I am glad to see roller camshafts in push rod engines, multiple valve overhead cam engines, turbo and super chargers used, and electronic engine management systems that allow each cylinder to be tuned separately. Production engines are far more technically advanced than the typical racing engine in this country. Racing engines will still have a mechanical valve train long after production engines are computer controlled. Fortunately, that is good for the camshaft industry. Let’s just match the proper cam profiles with the camshaft and the application, as it should be.
You hear the term “area under the lift curve“ used a lot in cam profile designs. Above is a very simple chart to help visualize what is meant by the area under the lift curve. I created a very basic chart with a drawing program. It does not represent any known cam profiles and is not to any scale or resolution. Its only purpose is a visual aid.
Remember the lift table from previous posts? It shows the profile lift for each degree of rotation with zero-degrees being the maximum lift. If you convert the lift table into a graph form, you have the lift curve. The horizontal line is the number of degrees in the profile and the vertical line is the profile lift. The chart only shows the opening side of the profiles. There are two cam profiles represented in the chart; both have the same lift and the same seat-to-seat durations. The blue is the area under the lift curve. Simple enough, right? There is obviously less blue area under the black line than there is under the blue line. Once the valve starts to open, there is more duration for the same lift on the blue profile. Instead of the seat-to-seat durations being the same, the two profiles could have the same duration at .050 or any other tappet height. This is a very meaningful way to compare cam profiles and to see the advantage of having more area under the lift curve.
If you came into this not knowing much about cam profiles, you should now have a better understanding having read all of my blog posts up to this point. If you do not understand something or have any questions, please do not hesitate to ask. Send me an email at firstname.lastname@example.org.
I usually get my ideas on topics from questions or from discussions on the forums I follow on the Internet. I try to post at least one entry during the month. If you are not seeing a new post, it means I need your help coming up with a new topic.
I am not going to write about the history of the internal combustion engine. There is already plenty written on the subject. For those that are interested in this early history and the beginnings of the motorized vehicle, I would do some reading on the following to get started:
De Dion-Bouton (company)
I enjoy reading about the early days and development of the internal combustion engine, the automobile, and the motorcycle. The above recommended topics will certainly open the door to an interesting time in history. In your studies, you will find that the camshaft did not play a very important role in the early engines.
Some Early Cam Profile Developers:
Pierre Louis Bertrand
I listened to a discussion the other day about maximum lift and centerlines. I honestly have never given it much thought and have always used the two terms interchangeably. After listening to the discussion closely and sifting through the facts, I came away with a different perspective.
When I talk about the intake or exhaust centerline, I am referring to the maximum lift point. On a symmetrical cam profile, they are one in the same. The confusion begins, when talking about an asymmetrical cam profile. The centerline and the maximum lift will be different on an asymmetrical profile. It all comes down to the definition of centerline.
One side of an asymmetrical cam profile has more total degrees than the other side. If zero degree is the maximum lift, then by definition, the centerline will not go through zero and the center point of the base circle. To divide the profile into equal opening and closing degrees, the centerline will need to be off from zero by half the amount of the asymmetry.
I do not have a picture, so I will try to explain this with words only. Drawing a picture would probably make this easier to understand. The opening side of a cam profile is 100- degrees. Maximum lift starts at zero and counts down the profile to the beginning of the opening ramp. This is the lift table, which has been talked about many times before. The closing side is 110-degrees. Maximum lift starts at zero and counts down the profile to the end of the closing ramp. The asymmetry of the profile is 10-degrees. A line drawn through the center of the base circle diameter and zero (maximum lift) would not divide the profile into equal degrees. The line would need to pass through a point 5-degrees to the left of zero (opening side) to be the true definition of centerline. The centerline would then divide the profile equally into 105-degrees on each side.
A camshaft grinder can use whatever method to create the specifications for his timing card. He can also make-up any method he chooses. The maximum lift is generally used for the reference point in my experience. The lobe separation angle on a camshaft is the angle between the intake and exhaust lobe maximum lift points. It only makes sense to use the maximum lift as the reference point for indexing the camshaft in the engine. When camshaft cores are made, the maximum lift is used to index the lobes with the dowel pin or keyway. When programming a cam profile into a CNC machine, the maximum lift is used as the reference point.
All of this is really no big deal. It is more of an argument over the true definition of centerline than anything else. I have always (and will continue to do so) used the maximum lift of the cam profile for a reference point. I have also unknowingly used the term centerline to mean the maximum lift. From now on, I will stop using the term “centerline”, and only use “maximum lift” in any camshaft discussions. That should eliminate any confusion.
The first thing that comes to mind when one thinks about a cam profile is the actual shape of the lobe. In the cam profile design world, the lift table is the cam profile, not the shape of it. The actual cam profile shape (excluding the physical dimensions) is of little importance.
As talked about in previous blogs, the lift table shows the tappet height at each degree of cam profile rotation. There is also a sample lift table shown in a previous blog. What may not be clear to some is that the lift table does not represent any particular type of tappet. The table only shows tappet height. The lift table could be exactly the same whether the tappet was a roller or a flat style. The table could also be the same no matter the diameter of the tappet or roller wheel. Sound confusing? It’s really not.
The cam profile shape is created by the lift table and the tappet used. As long as it is physically possible, any tappet can be used with any lift table. Using the same lift table, the type of tappet will determine the shape of the cam profile. Obviously, a roller profile and a flat profile will look different, but the valve movement will be exactly the same. The tappet height at each degree of rotation will be the same. The lobe shape combined with a particular tappet will determine the tappet height. Make sense now?
Many times a flat profile cannot be made from the same lift table as a roller profile, because the velocity will be too high for the diameter of the flat tappet face. A roller profile might encounter a severe negative radius when made from the lift table of a flat profile. Make sure you have read and understand all of the previous blogs if this is not making sense. If none of these or any other physical problems occurs, the same lift table can be used to produce a flat and a roller cam profile. The finished cam profile would still need to comply with all of the other design parameters to be a good profile.
To sum it up, the profile design procedure will create the lift table. Based on the type of tappet that will be used, the cam profile shape is then created and the final design is accepted after much trial and error until all of the parameters are met. Computer software allows all of this to happen very quickly even though separate steps are actually taking place.
Valve contact is the term I have always used to describe the point on the opening side of the cam profile where the valve first starts to move. On the closing side, this point is where the valve first makes contact with the seat. These points are on all cam profiles, both solid and hydraulic. I do not think this is an official cam design term, as I have heard other terms used.
The valve contact point is used when designing the cam profile ramps. On a solid cam profile, the valve lash will be determined by where the valve contact points are and the rocker arm ratio used. Some cam profiles are designed without much attention given to the ramps and the valve contact points. The valve contact and the lash is decided on after the cam profile is designed. I do not do this. I design the ramps first, based on the type of cam profile and the application. I then design the cam profile and mate it to the opening and closing ramps. On both a solid and a hydraulic cam profile, the valve contact points will determine the velocity at which the valve will open off and close on its seat. The proper location of the valve contact points will contribute to a smooth and reliable valve train for the application.
It has been a little over a year since I created this website and I am starting my fifth year in business next month. I just want to thank everyone that has supported me and those that have given me the opportunity to design your cam profiles.
I have created a simple, handy, spreadsheet program that will calculate the valve timing events for symmetrical and asymmetrical cam profiles. It will also give valve lift and valve lash data along with the minimum flat lifter diameter based on the maximum velocity. It can be downloaded from the engine equation page. If you would rather request the program, I can email it to you. It will run with any version of Microsoft Excel.
I would recommend reading the previous entries before starting here. I have intentionally not explained in detail what some things are, since I have already done so in the previous entries.
Every design method has the same limitations to deal with when designing a cam profile. These limitations will make it difficult to prove which method is better.
A flat tappet profile has a velocity and a nose radius limit. The tappet diameter alone will determine the maximum velocity that can be designed into the profile. The nose radius is determined by a combination of the base circle diameter, lift, and nose acceleration.
A roller tappet profile has a pressure angle and a negative radius limit. The pressure angle is the angle between where the lobe and roller wheel make contact. It is mostly determined by a combination of the base circle diameter, roller wheel diameter, and the lift. The negative radius is the concave area on the opening and/or closing flank of the profile. It is determined by a combination of the base circle diameter, lift, acceleration and roller wheel diameter.
Now you can see the importance of a large base circle diameter and roller wheel diameter when designing a cam profile. That is the reason for the trend to the larger camshaft journals. Strength is also a factor.
All of these limitations can cause premature wear and mechanical damage to the valve train if they are exceeded far enough.
Fortunately, there are some design "tricks" that can help deal with these limitations. On some flat tappet profiles, the maximum velocity can be dwelled (DMV) for a certain number of degrees, to increase the lift and duration without causing a sharp nose radius. This will not work for all flat tappet profiles. There are some profiles that just cannot be made. For example, design me a cam profile for the following: Small Block Chevy 1.8685 journal diameter, 0.842 OEM flat tappet, 240-degrees at 0.050 tappet height, 0.400 lobe lift. Sounds simple enough, right? Please send me the basic lift table when you are finished. You will become my new best friend. Thank you in advance. A technique can be applied to roller tappet profiles that will dwell the radius of curvature for a certain number of degrees. This will keep the maximum acceleration from going too high and creating too small of a negative radius. An acceleration curve chart will immediately reveal if one of these techniques was used to design the cam profile.
(a little secret) As long as there are limitations, all cam designers will eventually end up with basically the same cam profile design, for a given application. Some designers will just spend more time on their designs to make them the best that is possible within these limitations.
This is part 3 on this topic. Go back and start at part 1. Below is a typical cam profile lift table. The lift table is really what a cam profile is all about. The actual shape of the lobe will be determined by this lift table and the tappet that is used. This shows only the opening side of the lift table. The table shows the tappet height at that degree of lobe rotation. Zero degree is the maximum lift point. Ninety-four degrees is the beginning point. Sort of the opposite of what you would think, but this is the correct layout and will make sense later.
From the lift table, the velocity, acceleration, and jerk tables can be calculated. Each of the four tables can then be graphed to give a visual chart that is easier to look at and interpret. The acceleration graph is what gets the most attention. The acceleration curve chart on the home page was created from this lift table. All of the four tables and charts show valuable information, but the lift table is the one used to make the cam profile.
Like many things, people know how to do something, but do not understand why. I always want to know why. All of us know the common duration at 0.050 camshaft specification. It is accepted without much thought of why or where is really comes from. Do you know the duration at 0.050 from this lift table? If you look down the lift column, find the closest number to 0.050. It is 0.0502788. The degree related to this number is 66. That means at 66-degrees from maximum lift the tappet lift is 0.0502788. The exact degree for 0.050 would be 66.06 from linear interpolation. Now you see why the lift table layout is this way. All of the valve timing figures are based off of maximum lift. If the closing side of this profile were the same, 66.06 would also be the degree that 0.050 tappet lift takes place. If the closing side is different (asymmetrical), the degree may be some other number. Adding these opening and closing degrees at 0.050 tappet lift equals the duration in camshaft degrees. For crankshaft degrees, which is what is published, multiply camshaft degrees by two. Remember, the crankshaft rotates twice as fast as the camshaft. These numbers will also be used to calculate the opening and closing points for the valve timing. That is why for an asymmetrical cam profile, these numbers from the lift table must be known. Now you know why.
This is a continuation of the previous entry. Go back and read it first. There is a lot of talk about what design method is best to design a cam profile. You hear words like polynomials, splines, B-splines, NURBS, and circular arcs. I approach most things in a realistic way. Everything that man creates goes through a design phase, usually right after the idea is created. Sometimes the design looks good in the design world, but will not work in the real world. This can certainly apply to cam design.
When a modern cam profile is designed, the lift table is actually what is designed. The shape of the lobe is a product of the lift table and the tappet used. The lift table is simply the movement of the tappet compared to the rotation of the lobe. Usually displayed in 1-degree increments, the lift table will show the tappet lift per degree of rotation. A good display will show the lift out to seven or eight decimal places. Except for the precise numbers, really simple... right?, nothing complicated.
If your mind is capable, a lift table can be designed just from your brain. There are people that can actually do this. For real. No computer, no calculator involved. What do you call this cam design method? A cam profile designed this way is unique, because any numbers can be used, which could create a profile that no other method would be able to do.
In the beginning, the shape of the lobe was designed as the first step. No thought was given to a lift table. There was a base circle diameter and the maximum lobe lift with a nose radius. Two lines were drawn tangent to connect the base circle with the nose radius and tada; you have a cam profile. In this case, the profile shape would create the lift table. Make sense so far. When people started to analyze the movement of the tappet from the profile shape, the lift table was created. It was far from acceptable compared to today's cam profile designs. People in the know soon realized that by controlling the tappet movement, more power was made by the engine. We are just talking about tappet movement, nothing about opening and closing, centerlines or lobe separations. That is camshaft design, which comes after cam profile design.
Through the years many techniques, formulas, programs, and methods have been used to design cam profiles. Some will create a smoother profile than others. Some are easier to use than others. When a cam profile is created, it has specifications that the designer is trying to meet. Important to most people are the lift and durations of the profile. If the profile was "maxed-out" (my terminology) when designed, the velocity, acceleration, and jerk curves are all at maximum levels, but will still produce a smooth, reliable profile. The duration figures at different tappet heights will also be at maximum values. In order to get more duration (more area), the other maximum levels will have to be exceeded.
Whatever method is used to design a cam profile, the real world will cause each method to reach the same limitations. Many designers will focus on the smoothness of the profile and pick a particular method because it does indeed smooth-out the profile. It looks good on paper or the computer screen, but more area under the lift curve is not increased. This area is what makes power. That is the reason for the larger base circle diameters and lifter diameters. They will allow more area to be created without increasing the other important limits. Software alone will not create more area without exceeding the limits. If the area did not matter, we could have just kept creating a shape and not worried about how fast to raise and lower the tappet.
Like most things, time will tell. I am always following and studying this subject. The method of manipulating the numbers in the lift table looks good to me. A program would need to be created that would do this and then display the velocity, acceleration, and jerk curves. Maybe allow the user to input limits and ranges. That would be cool.
A cam profile is nothing more than a mechanical device that uses a rotational motion in order to create a lifting motion in something. Similar to a lever, but instead uses a rotational motion to create lift, whereas a lever uses a linear motion.
Unrelated to cam profiles for camshafts, there are spiral cams that use a spinning motion to create lift. A screw type bottle jack would be one example of this type of cam.
A cam profile used for camshafts will exert the lifting motion onto a tappet. Lifter and follower are sometimes used to describe the same part. Picture the cam profile and the tappet in contact with each other on a vertical axis. Now, imagine rotating the cam profile and lifting the tappet. The cam profile lifts the tappet to maximum lift and then brings the tappet back down. A very simple process, right?
Simple, but when a cam profile is designed, certain parameters must be observed. If the profile is divided into 360 degrees, then we can observe what is happening at each degree of rotation. The first thing you will notice is how much the tappet moved (lifted) for each degree the profile was rotated. The lift per degree of rotation varies as the profile is rotated. It will begin with very little lift per degree and increase until the maximum lift per degree is reached. The tappet lift per degree of profile rotation is called the velocity and is expressed in inches per degree.
The velocity is important in the design of a flat tappet profile. The diameter of the flat tappet will determine the maximum velocity that can be designed into the profile. If the maximum velocity is exceeded, the contact point will run off the edge of the tappet and cause immediate damage. Roller tappets obviously do not have this problem.
Some other common parameters are acceleration and jerk. Acceleration is the rate of change in velocity and jerk is the rate of change in acceleration. Both are important in the design of cam profiles. Years of experience will determine these values.
Another important value for a flat tappet profile is what's called the nose radius. This determines the shape of the nose of the profile. Ideally, you want the profile to have a broad rounded nose. If the nose radius is too sharp then premature wear will take place. On a roller profile, there is what's called a negative radius. This is when the radius is inverted. On one or both sides of the lobe profile (the flank), the shape will go slightly inward and then back out again creating a negative radius in that area. If the negative radius is small enough, a special setup with a smaller wheel is necessary to grind the profile shape. This is labor intensive and will cost more. This type of profile is also harsh on the valve train requiring more maintenance. Again, years of experience will determine the nose radius and how much negative radius is acceptable.
To sum it up, the mechanical operation of a cam profile is fairly simple; to properly design a cam profile is a little more complicated.
The following is one of the better articles I have found on camshafts. It is written with Harley-Davidson motorcycles in mind, but certainly applies to automobiles as well. Very informative...enjoy.
What you need to know about Cams.
#1 - the most common camshaft error made by people is to OVER cam the engine.
#2 - is selecting a cam that is not compatible for the RPM range that we plan to operate the engine in.
There are a number of things, which affect the cam’s performance.
Cylinder Head Flow Rates: The cylinder heads must be able to flow enough air for the time that the valves are open.
Compression Ratio: Static compression ratio and cam choice should be considered as a system.
Intake: The intake must also flow enough air to support the cam and cylinder fill.
Exhaust Pipes: The exhaust pipe must not only flow enough, they must be designed so that the reversion pulse is compatible with camshaft timing.
There is a little clue here for you sharpies and that is the word AIRFLOW. Airflow is everything and the camshaft is the controller of the airflow. It determines how much, when and how long. The result of all the camshaft specifications is where in the RPM band the motor will make the best power. Now one more thing before we dive in to the mystery and that is we need to understand our objective here. For the purposes of these articles I'm not interested in a RACING motor but rather street motors and street motors need to develop TORQUE and they need to develop TORQUE in the 2500 to 4500 RPM range, as this is the range we most often operate the engine in (freeway cruising). There is a law of camshafts we need to keep in mind. If you have it at the top, you will not at the bottom and if you have it at the bottom, you will not at the top. We cannot have it all. In a STREET motor, we do not need it at the top, we do need some at the bottom, but we really need mid-range. So here is where we are going to look for our torque. Now we also have to look at the bike we ride. A dresser is going to need more bottom end than a FXR due to the weight and wind resistance. All right, so now that we have all this behind us we can move on to the cam itself.
Intake Closing: The intake closing point has more effect on engine-operating characteristics than any of the other three opening and closing points. The earlier it occurs, the greater the cranking pressure. Early intake closing is critical for low-end torque and responsiveness and provides a broad power curve. It also reduces exhaust emissions while enhancing fuel economy. As RPM increases, intake charge momentum increases. This results in the intake charge continuing to flow into the combustion chamber against the rising far past BDC. The higher the engine's operating RPM, the later the intake closing should be to ensure all the charge possible makes it into the combustion chamber. Of course, closing the valve too late will create significant reversion. It is a fine balancing act. In a perfect world, the optimum intake closing point would occur just as the air stops flowing into the chamber. It would get the valve seated quickly and not waste time in the low lift regions where airflow is minimal and there is no compression building in the cylinder. It wouldn't be so fast the valve bounces as it closes, allowing the charge to escape back into the intake port and disturb the next charge. And in hydraulic street cam applications, it would insure that the closing ramps are not so fast that they result in noisy operation.
A late closing intake valve will yield poor compression and will cause poor performance over most of the entire RPM range.
A semi-late closing intake will have a good mid range and good top end but not the best.
An early closing intake (30-35 degrees) is what we like for a heavy bike because it will give an excellent bottom end performance and a good mid-range.
The intake valve closing point is intimately related to an engine's dynamic or "effective" compression ratio. Compression ratio is also dependent on cam duration.
A mild cam with an early intake valve closing point will work well at low RPM. However, at high RPM the intake valve will close before the maximum amount of air/fuel mixture has been drawn into the cylinder. As a result, performance at high RPM will suffer. If a high static compression ratio is used with a mild cam (i.e. and early intake valve closing point) then the mixture may end up being "over-compressed.” This will lead to excessive compression losses, detonation and could even lead to head gasket or piston failure.
On the other hand, an aggressive cam with a late intake valve closing point will work well at high RPM. However, at low RPM the intake valve will close too late for sufficient compression of the intake charge to occur. As a result, torque and performance will suffer. If a low static compression ratio is used with an aggressive cam (i.e. a late intake valve closing point) then the mixture may end up being "under-compressed.” Thus, a high performance cam with long duration should ideally be combined with a higher static compression ratio. That way the engine can benefit at high RPM from the maximized amount of intake charge afforded by the late intake valve closing, and still achieve sufficient compression of the mixture as a by-product of the dynamic compression ratio.
Intake Closing AGAIN!: The most important cam event. This sets the engine's effective RPM range, effective dynamic compression. An early closing (30 - 38 ABDC) = high dynamic compression, great low to mid RPM torque for a very broad power band, requires lower static compression (which means less stress and strain on the engine, less risk of heat damage and detonation, more reliability)....but engine RPMs are limited, the engine will "quit pulling" around 4800 RPM. As intake valve closing gets later (40-45) the power band moves up about 250 -300 RPM, narrows slightly unless more static compression is built in (e.g. thinner head gasket). Torque remains about the same, but due to higher RPMs, HP increases slightly. Throttle response off of idle drops slightly. Head temperatures increase slightly, making detonation a realistic risk, fuel management/tuning becomes even more critical. And exhaust pipe diameter, length, back-pressure designs become more influential. The engine will pull thru 5000 RPM. Closing the valve even later (+45 ABDC) shift the power band way up the RPM scale. Increased static compression is necessary to achieve any TQ/HP. Typically, it will exceed 12:1. Fuel management/tuning are very critical to reduce detonation and the risk of heat damage. Higher compression shortens the engine's life. Because this cam only functions well at higher RPMs, the other cam specifications can take advantage of this and be optimized for more power. What's lost is smooth idling and some usable power/torque at low to mid RPMs, crisp throttle responses from idle, engine heat issues become critical. Think about a bad-ass quarter mile drag bike: won't idle for crap, pops & snorts until the throttle is twisted nearly WFO, when it finally begins to roar - the engine is barely manageable - but damn, what a ride!
The intake closing point has more effect on engine-operating characteristics than any of the other three opening and closing points. The earlier it occurs, the greater the cranking pressure. Early intake closing is critical for low end torque and responsiveness and provides a broad power curve. It also reduces exhaust emissions while enhancing fuel economy. As RPM increases, intake charge momentum increases. This results in the intake charge continuing to flow into the combustion chamber against the rising far past BDC. The higher the engine's operating RPM, the later the intake closing should be to ensure all the charge possible makes it into the combustion chamber. Of course, closing the valve too late will create significant reversion. It's a fine balancing act. In a perfect world, the optimum intake closing point would occur just as the air stops flowing into the chamber. It would get the valve seated quickly and not waste time in the low lift regions where airflow is minimal and there is no compression building in the cylinder. It wouldn't be so fast the valve bounces as it closes, allowing the charge to escape back into the intake port and disturb the next charge; and in hydraulic street cam applications, it would insure that the closing ramps are not so fast that they result in noisy operation.
The most important timing event is the intake valve closing angle. The intake closing point determines the minimum RPM at which the engine begins to do its best work. The later the intake valves close, the higher the RPM must be before the engine gets "on the cam."
If you have one of the late closing cam designs installed, say one that closes the intake valves later than 40 degrees, then you cannot expect excellent performance at 2000 RPM. No carburetor adjustment, ignition adjustment or exhaust system can change this.
Intake Opening: Looking at the intake valve, its opening point is critical to vacuum, throttle response, emissions, and gas mileage. At low speeds, and high vacuum conditions, premature intake opening during the exhaust stroke can allow exhaust gas reversion back into the intake manifold, hurting the intake pulse velocity, and contaminating the fresh intake charge. A late opening intake gives smooth engine operation at idle and low RPM, and it ensures adequate manifold vacuum for proper accessory operation (assuming the other three valve opening and closing points remain reasonable). As RPM increase, air demand is greater. To supply additional air and fuel, designers open the intake valve sooner, which allows more time for the intake charge to fill the cylinder. With an early opening intake valve, at high RPM, the exiting exhaust gas also helps draw the intake charge thru the combustion chamber and out the exhaust - that's good for purging the cylinder of residual gas, but it also increases fuel consumption by allowing part of the intake charge to escape before combustion and can make for a rough idle.
Early usually means overlap, less throttle response at low to mid RPMs, rough idle, more emissions, poor fuel economy. However, by opening the intake valve early, we can slightly increase the volumetric efficiency of the engine...if the heads will flow any better. This is where stock heads fall short compared with ported heads. However - like cams, bigger is not always better when it comes to ported heads. Big ports and big valves will drop the intake and exhaust velocities, which can cause a host of problems, as well as a loss of volumetric efficiency. Most ported Twin Cam heads with stock diameter valves/seats, used with stock intakes and SE or K&N air filters usually flow their max CFM near 0.350" - 0.450" valve lift. Using a cam that has the intake valve open that far by the time the piston is at max velocity maintains the max intake charge velocity - which makes the best use of the momentum supercharging effect between idle and 3500 RPM. Using a cam with even more lift (+0.500") only reduces this effect - and power (along with adding more unnecessary wear and tear to the valve train). A stock, un-ported head has a very restrictive exhaust port, and therefore limits volumetric efficiency even further - making a cam with high lift even less effective. The thing to remember with cam timing is that the intake valve opens before TDC and closes after BDC.
Exhaust Opening: Overall, the exhaust valve opening point has the least effect on engine performance of any of the four opening and closing points. Opening the exhaust valve to early decreases torque by bleeding off cylinder pressure from the combustion that is used to push the piston down. Yet the exhaust has to open early enough to provide enough time to properly scavenge the cylinder. An early opening exhaust valve may benefit scavenging on high-RPM engines because most useful cylinder pressure is used up anyway by the time the piston hits 90-degrees before BDC on the power stroke.
Later exhaust valve opening helps low RPM performance by keeping pressure on the piston longer, and it reduces emissions.
Early opening exhaust - here we loose our entire bottom end and our mid range will be lazy what it will do is run hard on the top.
Semi-early opening exhaust. This timing will give us good cylinder scavenging which results in a cleaner cylinder mixture at high RPM the low end will suffer some but the mid range will be very good.
Late closing exhaust here we end up with a narrow RPM band the low end will be good as well as the mid-range but we will have an engine difficult to use.
Stock cams typically open the exhaust valve late (36 BBDC) to maximize the burn time and pass emission tests easier...but suffer from pumping losses because the piston has to work harder to mechanically push out the burnt gases. If the cam opens the exhaust valve a little sooner (40-43 BBDC), we can use blow-down (the expansion of burning A/F) to help scavenge the cylinder. This gets the burnt gases moving, reduces the piston effort, and decreases pumping losses...up too about 4000 RPM. However if the cam opens the exhaust valve too soon ( 45+ BBDC) the blow-down will bleed off much of the expansion pressure of the power stroke from idle thru about 2500 RPMs. The RPMs must be higher to overcome the time available for blow-down.
Exhaust Closing: Excessively late exhaust valve closing is similar to opening the intake too soon- it leads to increased overlap, allowing either reversion back up the intake, or the intake mixture to keep right on going out the exhaust. On the other hand, late closing events can help purge spent gasses from the combustion chamber and provide more vacuum signal to the intake at high RPM. Early exhaust valve closing yields a smoother operating engine. It does not necessarily hurt the top-end, particularly if it is combined with a later intake valve opening. As engine operating range increases, designers must move all the opening and closing points out to achieve earlier openings and later closings, or design a more aggressive profile to provide increased area under the curve without seat timing increases. Exhaust Valve Closing - usually between 4 (early) and 20 (late) deg ATDC. An early closing = less overlap, late closing = large overlap. Less overlap (exhaust valve closes at 4) makes it easier to pass a smog test, smooth idle, great fuel economy. A mild overlap (exhaust valve closes at 8-12) makes good low to mid RPM range power, better throttle response, fair fuel economy, slightly more emissions. And large overlap (exhaust valve closes at 13-20) allows a lot of intake charge dilution/loss (bad emissions), poorer fuel economy, rough idle, less throttle response from idle, and makes most of the power at higher RPMs. Note: the amount of overlap also depends on the cam's intake valve opening specifications.
Early exhaust valve closing yields a smoother operating engine. It does not necessarily hurt the top-end, particularly if it's combined with a later intake valve opening. As engine operating range increases, designers must move all the opening and closing points out to achieve earlier openings and later closings, or design a more aggressive profile to provide increased area under the curve without seat timing increases.
Lobe Centerline: Lobe Centerlines give you a relative perspective of how advanced or retarded a cam is in relation to top dead center (TDC). Harley cam profiles typically have an intake centerline from 98 to 108 degrees. An intake centerline of 98 is considered to be the most advanced and generally gives the most torque. A centerline of 108 will give power in the upper RPM range.
An exhaust centerline of 112 is the most advanced while the 102 is the most retarded. Again an advanced lobe will give power in the lower RPM range while the retarded lobe will have its power range extended in the RPM range. For practical terms, most cams for Harley are in the range of 96-108 on intake and 112-102 on the exhaust.
Tailoring the valve opening and closing points on an actual camshaft is accomplished by varying the lobe centerline locations, changing the LSA, and refining the profile shape itself. Advancing the cam moves both the intake and exhaust in an equal amount, resulting in earlier valve timing events. Engines typically respond better with a few degrees of advance, probably due to the importance of the intake closing point on performance. For racing, advanced cams benefit torque converter stall, improve off-the-line drag race launches, and help circle-track cars come off the corner. Cam companies often grind their street cams advanced (4 degrees is typical), which allows the end-user to receive the benefits of increased cylinder pressure yet still install the cam using the standard timing marks. Increasing the intake lobe centerline from 104 to 106 degrees is considered retarding. All events will take place later in the engine cycle. Retarding the cam causes the intake valve to open and close later. This will reduce cylinder pressure which reduces the low speed performance of the engine.
Advancing the intake and retarding the exhaust (“closing up the centers”) increases overlap and should move the power up in the RPM range, usually at the sacrifice of bottom end power. The result would be lower numerical values on both intake and exhaust lobe centers.
Retarding the intake and advancing the exhaust (“spreading the centers”) decreases overlap and should result in a wider power band at the sacrifice of some top end power. This condition would be indicated by higher numerical values on both intake and exhaust lobe centers. By moving only one cam the results are less predictable, but usually it is the intake that is moved to change power characteristics since small changes here seem to have a greater effect.
Lobe Separation Angle: Lobe separation is the angle between the center bump of the intake lobe and its counterpart on the exhaust lobe. Think of it like the two points on a pair of scissors relative to the hinge in the middle. If the scissors are nearly closed, you can cut well as long as what you are cutting is thin. To cut thick stuff, you open wider, but have less leverage, so it can be harder to get the done. The same principle applies with separation on cam lobes. Typically, lobe separation for street cams runs between 97 and 108 (camshaft) degrees. The relationship between intake and exhaust is ground into the cam and can’t be altered by advancing or retarding the overall cam timing.
As a guideline, if the rest of the numbers are comparable, a cam with a lobe that is less separate (for example, 98 to 103 degrees) will offer a broader spread of power and tend to produce power at the low end, while wide lobes make for a more “cammy” cam, coming on harder and later in the game. Lobe Separation Angles (LSA) of 100-103 degrees tends to produce power at the low end.
LSA and Lift affect the "sound" and idle quality. Generally, smaller lobe separation angles cause an engine to produce more mid-range torque and high RPM power, and be more responsive, while larger lobe separation angles result in broader torque, improved idle characteristics, and more peak horsepower.
A “tight” lobe separation angle of 103 degrees or less creates more valve overlap, which helps create that lumpy idle characteristic of big camshafts. The tighter LSA’s are, the more likely problematical exhaust reversion into the intake will occur. Put simply, we can say that a tight LSA cam produces a power curve that is, for want of a better description, more "punchy." At low RPM when off the cam, it runs rougher, and it comes on the cam with more of a "bang." Narrow LSA’s tend to increase mid-range torque and result in faster revving engines. Generally, smaller lobe separation angles cause an engine to produce more mid-range torque and high RPM power, and is more responsive. Typically, however, small lobe center numbers (more overlap) equates to more mid-range power at the expense of top-end power. Probably the most significant factor to the engine tuner though is a tight LSA’s intolerance of exhaust system back-pressure. Remember, during the overlap period both valves are open. If there’s any exhaust back-pressure or if the exhaust port velocities are too low it will encourage exhaust reversion. A cam with 102 degrees of lobe separation angle will have more overlap and a rougher idle than one with 108 degrees, but it'll usually make more mid-range power. A tighter lobe has more overlap. A tighter centerline starts torque curve sooner, and doesn't give a wide power band. A wider lobe doesn't start the torque curve sooner, but it continues to make torque longer and has a broader power band.
Wide LSA’s result in wider power bands and more peak torque at the price of somewhat lazier initial response. Larger lobe separation angles result in broader torque, improved idle characteristics, and more peak horsepower. A wider lobe doesn't start the torque curve sooner, but it continues to make torque longer and has a broader power band. A street engine with a wide LSA has higher vacuum and a smoother idle. Big numbers (less overlap) will give more top end, sacrificing mid-range. A cam on wide centerlines produces a wider power band. It will idle smoother and produce better vacuum, but the price paid is a reduction in output throughout the working RPM range.
Narrow LSA (98-103)
Moves Torque to Lower RPM
Increase mid-range Torque
Increases Maximum Torque
Faster revving engine and more responsive
Narrow Power band
Builds Higher Cylinder Pressure
Increase Chance of Engine Knock
Increase Cranking Compression
Increase Effective Compression
Idle Vacuum is Reduced
Idle Quality Suffers (lumpy idle characteristic)
Open Valve-Overlap Increases
Closed Valve-Overlap Increases
Decreases Piston-to-Valve Clearance
Wide LSA (104-108)
Raise Torque to Higher RPM
Reduces Maximum Torque
Broadens Power Band
Lazier initial response
More peak Horsepower
Reduce Maximum Cylinder Pressure
Decrease Chance of Engine Knock
Decrease Cranking Compression
Decrease Effective Compression
Idle Vacuum is Increased
Idle Quality Improves
Open Valve-Overlap Decreases
Closed Valve-Overlap Decreases
Increases Piston-to-Valve Clearance
Overlap: The objective of overlap is for the exhaust gas which is already running down the exhaust pipe, to create an effect like a siphon and pull a fresh mixture into the combustion chamber. Otherwise, a small amount of burned gasses would remain in the combustion chamber and dilute the incoming mixture on the intake stroke. Duration, lift and LSA combine to produce an "overlap triangle". The greater the duration and lift, the more overlap area, LSA’s remaining equal. Given the same duration, LSA and overlap are inversely proportional: Increased LSA decreases overlap (and visa versa). More overlap decreases low RPM vacuum and response, but in the mid-range, overlap improves the signal provided by the fast moving exhaust to the incoming intake charge. This increased signal typically provides a noticeable engine acceleration improvement.
Less overlap increases efficiency by reducing the amount of raw fuel that escapes thru the exhaust, while improving low-end response due to less reversion of the exhaust gasses back up the intake port; the result is better idle, stronger vacuum signal and improved fuel economy. Due to the differences in cylinder head, intake and exhaust configuration, different engine combos are extremely sensitive to the camshaft’s overlap region. Not only is the duration and area of the overlap important but also its overall shape. Much recent progress in cam design has been due to careful tailoring of the shape of the overlap triangle. According to Comp Cams, the most critical engine factors for optimizing overlap include intake system efficiency, exhaust system efficiency, and how well the heads flow from the intake toward the exhaust with both valves slightly open.
Camshaft overlap duration less than 30 degrees tends to produce good low end power.
Overlap amounts to the time the intake and the exhaust valves are both open. When you get it right, overlap helps draw in the intake charge, but excessive amounts actually reduce power by letting intake charge escape out the open exhaust valve. Lots of overlap virtually guarantees a cam won’t work well at low RPM, regardless of how strong it is wide open. Camshaft overlap duration of less than 30 degrees tends to produce good low end power.
Increased overlap equates to reduced idle quality, vacuum, and harsher running prior to coming up on the cam. Lots of overlap works great at high RPM because more intake charge manages to cram itself into the cylinder, but lots of overlap will also make the engine run badly at low RPM, as exhaust gas manages to make its way back up the intake manifold, diluting the incoming air/fuel charge, and depositing soot on the intake runners, carburetor, etc. Cams with a lot of overlap tend to cause rougher idling because of the lack of vacuum they create in the manifold.
Overlap (lots of duration and tight lobe-separation angles) decreases cylinder pressure, especially at low RPM, which allows an engine to run a higher compression ratio and still work on pump gas. High cylinder pressure, which is caused partly by a high compression ratio, is what makes an engine detonate on pump gas. Decreasing the cylinder pressure by adding duration is just like taking compression out of the engine, but mostly only at low RPM.
Duration: Duration has a marked affect on a cam's power band and drive-ability. Higher durations increase the top-end at the expense of the low end. A cam's "advertised duration" has been a popular sales tool, but to compare two different cams using these numbers is dicey because there's no set tappet rise for measuring advertised duration. Measuring duration at 0.053-inch tappet lift has become standard with most high-performance cams. Most engine builders feel that 0.053” duration is closely related to the RPM range where the engine makes its best power. When comparing two cams, if both profiles rate the advertised duration at the same lift, the cam with the shorter advertised duration in comparison to the 0.053” duration has a more aggressive ramp. Providing it maintains stable valve motion, the aggressive profile yields better vacuum, increased responsiveness, a broader torque range, and drivability improvements because it effectively has the opening and closing points of a smaller cam combined with the area under the lift curve of a larger cam. Engines with significant airflow or compression restrictions like aggressive profiles. This is due to the increased signal that gets more of the charge through the restriction and/or the decreased seat timing that results in earlier intake closing and more cylinder pressure. Big cams with more duration and overlap allow octane-limited engines to run higher compression without detonating in the low to mid-range. Conversely, running too big a cam, with too low a compression ratio leads to a sluggish response below 3,000 RPM. Follow the cam grinder’s recommendations on proper cam profile-to-compression ratio match-up.
Duration generally ranges from 220 degrees for a torquey bottom-end cam all the way to 295 degrees for a “top end rush,” typically measured at 0.053 inch lift.
As a general rule, lower-duration cams in the neighborhood of 210 to 200 degrees at 0.053 work best for stock-type replacement cams. Stepping past 220 degrees of duration (at 0.053) places the cam into the bolt-on, mid-range style category. These cams work well with the stock compression, intake and exhaust. Cams with 240-plus degrees of duration or more are beginning to step into the performance arena and generally work better with other induction, compression, and exhaust modifications. Duration has a marked affect on the cams power band and drivability.
Higher durations increase the top-end at the expense of the low end. As a general rule, cams with 220-235 degrees of duration tend to produce good low end torque. Cams with 235-250 degrees of duration tend to work best in the mid-ranges and cams over 260 degrees work best for top end power.
It is important to remember here that the duration values given are to be used as a general rule and that increasing the duration will have an effect on the idle characteristics and overall drivability.
Long duration, late intake closing cam designs are necessary to drag the last bit of power out of an engine. Unfortunately, these same cams can perform poorly under more normal riding conditions. In the quest for maximum power output, many-too-many Harley owners choose a late closing, high-RPM cam for their engine. The problem with such choices is that the engine seldom spends time in the RPM range favored by such cams.
Lift: Another method of improving cam performance is increasing the amount of lobe lift. Designing a cam profile with more lobe lift results in increased duration in the high-lift regions where cylinder heads flow the most air. Short duration cams with relatively high lift can provide excellent responsiveness, great torque, and good power. But high lift cams are less dependable. You need the right valve springs to handle the increased lift, and the heads must be set up to accommodate the extra lift. There are a few examples where increased lift won't improve performance due to decreased velocity through the port; these typically occur in the race engine world (0.650- to 1.00-inch valve lift). Some late model engines with restrictive throttle-body, intake, cylinder head runner and exhaust flow simply can't flow enough air to support higher lift.
Cam (or lobe) lift is the maximum height or distance that the lifter or follower is raised off the cam. More lift generally means better top-end power, but you’ll sacrifice bottom-end response. In addition, cams with high lift typical put more wear and tear on the valve train.
For street bikes, lift figures are best kept at or below 0.500 inch, simply because, with the right cam, you can still get all the power you can use, but you won’t need a new valve train every 20,000 miles. Sure, with the right cylinder head/piston combination, lifts in the mid 0.500 inch range, even perhaps encroaching on 0.600 inch can work, but pushrods flex, geometry goes AWOL, and the extra benefits of the lift are quashed by the limits of flow through the ports (particularly the exhaust port), so why bother? Mega lift is more valuable to drag racers who re-engineer the whole plot any way.
The other potential problem with increasing the cam lift is that there is only so much clearance between the piston and the valve. The other problem associated with elevated lift numbers is spring fatigue. The greater the lift the farther the spring will have to expand and contract during each rotation of the cam. Cams with more lift are much harder on springs, causing a reduction in spring life.
Symmetrical Cams: This simply means that the cam lobe is the same on both sides. This means that the valve opens and closes at the same rate.
Asymmetric Lobes: In the past, both opening and closing sides of a cam lobe were identical. Most recently, designers developed asymmetrical lobes, wherein the shape of the opening and closing sides differ. Asymmetry helps optimize the dynamics of a valve train system by producing a lobe with the shortest seat timing and the most area. The designer wants to open the valve as fast as possible without overcoming the spring's ability to absorb the valve train's kinetic energy, and then close the valve as fast as possible without resulting in valve bounce. There are many different theories about how to design the most aggressive, stable profile. Hydraulic lifters can provide quiet valve train operation only if the closing velocity is kept below a certain threshold. However, the opening velocity can be higher and still provide quiet operation. Almost all modern hydraulic profiles have some symmetry.
Here the lobes differ from the opening side to the closing side. This allows the cam grinder to open the valve a one speed and close it at another. Here is where some cams are quite and some noisy. If the grinder has chosen to set the valve down slowly on the seat it will be a quitter cam than if the grind lets the valve down too quickly. Single pattern cams In the case of single pattern cams both the intake and exhaust lobe are the same. A cam can be asymmetrical and single pattern or symmetrical and single pattern. Dual pattern cams have different profiles on the intake and exhaust lobes. A cam of this type can be any combination of asymmetrical or symmetrical of profiles.
Camshaft Noise: Camshaft noise is partly from cam shaft ramp design and partly mechanical noise from end play and excess gear lash. Camshaft noise and gear lash is dictated by the cam support plate. When the teeth of the gears mesh, they produce annoying whine if they mesh too tightly and a clackety clatter if they’re too loose. However, these gears also expand slightly when the engine is at operating temperature and then return to their original size when the engine cools down, which is why it’s impossible to get them to be quiet all the time.. Throw in loose manufacturing tolerances on the cam support plate and you have a complicated issue on your hands, which is why Harley replaced gear driven cams with chain driven cams.
Effect of Compression Ratio on Camshaft Selection: It is instructive to remember that the static compression ratio that your engine displays on paper does not translate directly to higher cylinder pressures. The cylinder pressure (prior to ignition) during engine operation is dependent on what can loosely be called "dynamic or effective compression ratio". The pressure is greatly affected by the timing of your valve events - i.e. cam duration and timing. Specifically, the intake valve closing point is intimately related to an engine's dynamic or "effective" compression ratio.
But we just learned on static compression ratio is directly related to stroke. In principle, the piston cannot compress the mixture until the intake valve closes. Thus if the intake valve closes when the piston has already moved quite some distance up the bore, then the amount that the intake charge will be compressed is reduced. The "effective compression stroke" has been reduced. Does this mean that when an engine is operating that the dynamic compression ratio is lower than the static compression ratio? Well yes and no.
An engine with a performance cam operating at low RPM will suffer a loss of torque due to the fact that the effective compression ratio is reduced by the late intake valve closing point. However, as the RPM increases "inertia supercharging" becomes important. At high RPM's the intake charge is moving into the cylinder at high velocity. As such it has a lot of inertia and will continue moving into the cylinder past BDC, even though the piston has changed direction and is now moving up the bore (towards the incoming charge). Ideally the intake valve will close just before the incoming air stops and reverses direction. This guarantees that the maximum amount of air/fuel mixture has been drawn into the cylinder prior to ignition. When this happens an engine is said to have "come on the cam". In order to ensure that the mixture is still compressed sufficiently over the reduced effective compression stroke it is necessary to increase the static compression ratio. This is why high performance engines with aggressive camshafts also tend to have high static compression ratios.
Bottom Line: Static compression ratio and cam choice should be considered as a system.
A mild cam with an early intake valve closing point will work well at low RPM. But at high RPM the intake valve will close before the maximum amount of air/fuel mixture has been drawn into the cylinder. As a result performance at high RPM will suffer. If a high static compression ratio is used with a mild cam (i.e. and early intake valve closing point) then the mixture may end up being "over-compressed". This will lead to excessive compression losses, detonation and could even lead to head gasket or piston failure.
On the other hand, an aggressive cam with a late intake valve closing point will work well at high RPM. But at low RPM the intake valve will close too late for sufficient compression of the intake charge to occur. As a result torque and performance will suffer. If a low static compression ratio is used with an aggressive cam (i.e. a late intake valve closing point) then the mixture may end up being "under-compressed". Thus a high performance cam with long duration should ideally be combined with a higher static compression ratio. That way the engine can benefit at high RPM from the maximized amount of intake charge afforded by the late intake valve closing, and still achieve sufficient compression of the mixture as a by-product of the dynamic compression ratio.
Remember, questions and comments are always welcome.
The newest approach (which really isn't new) in roller camshafts is to use a larger journal diameter and/or a larger roller wheel diameter. In order to get the performance gains, the cam profiles must be designed for these diameters. Many times they are not.
Years ago when flat tappet mushroom lifters were first being used, cam grinders would often times use profiles that were not designed for the larger lifter diameter. Sometimes the engine builder would install the larger diameter lifters without changing the camshaft. In reality when this was done, there was no difference in performance, other than what was perceived by the racer. Many times things are wrongly done because lack of knowledge and are not intentional.
Today, the same thing is being done with roller camshafts. The cam profiles on these camshafts are not designed for the base circle diameter of the lobe or the roller wheel diameter that is being used. Unless you are being assured by your camshaft grinder that the profiles were indeed designed for the new larger journal or larger roller wheel diameters, chances are the profiles were originally designed for a standard size lobe and roller wheel. Just as with everything, the consumer must be knowledgeable and ask questions.
When a different size lobe or roller wheel diameter is used, the contact point between the lobe and the wheel changes. This changes the velocity, acceleration, jerk, radius of curvature, and the pressure angle from what the original design was. Sometimes no mechanical harm is done, but sometimes mechanical problems do occur. Premature wear to the roller lifter and valve spring problems are the most common problems. Even with no mechanical problems, the performance is not up to where it should be. Just because something works or is adequate doesn't mean that the process is correct. The best engine power is made when a combination of components are designed to work together.
No matter what anyone wants to tell you, the difference between one properly designed profile and another is the area under the lift curve. Take the following example of two profiles being compared and they both have the same lobe lift and the same duration at 0.050. The profile with more duration at 0.100, 0.200, 0.300, and so on would be the better profile. This profile would have more area under the lift curve. Very seldom will you see these higher lift duration figures published unless the cam grinder is very confident in his profiles. The lash setting will also play a part when comparing profiles. A wider lash setting will use up more of the area than a narrow setting. The cam profile and the proper lash setting will tell the whole truth. Remember, I started out by saying properly designed profiles. A profile can be designed with more duration at the higher lifts (more area), but the profile may be too aggressive for the application, and can actually damage the valve train. My approach is to balance more lift area with smoothness and reliability of the valve train.
Are you having a difficult time finding the cam profiles that you want? Do you have a specific lift rule or duration rule at the race track that you run on? Maybe you have a particular cam profile in mind but no one offers it. Trying to get a camshaft grinder to design a cam profile just for you is almost impossible. It is just not cost effective for them to put in the time to design a special profile that may not be a big seller. Most would rather sell the profiles they already have. If they were to design your cam profile, the cost would probably scare you away.
Let me know what cam specifications you are looking for and I will gladly give you the time. Sometimes certain cam specs will not produce a good profile. I will be able to tell you if your cam specs will work or not. We can also discuss altering the specs if that is necessary.
If you decide to have me design your cam profiles, you can then send the design data to the camshaft grinder you choose, and have a camshaft made with the profiles that you wanted.