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Technical Advices > Valve Train and Camshafts

Valve Train and Camshafts

4.1 Valve Train Friction

The valve train is one of those almost forgotten items in a race motor. Like most other things, there is lots of folklore about what is good and what is not. Over the years I can count the amount of time that I have spent playing around with valve train components not in hours or days, but more likely in months.

  • First, it is important to understand that the valve train assembly consists of everything from the cam bearings to the valve seat, and everything in between.
  • Second, the valve train also accounts for the highest percentage of total parasitic loss within an engine. Whatever can be done to reduce parasitic loss WILL make a difference, albeit small at times.

Cam Bearings - Let’s start with the cam bearings. As Scuderia Topolino does on great deal of work on A112 type motors, I will focus on this type of installation, but the general ideas are applicable to almost any internal combustion engine. There are basically three cam bearings. The front one is pre-sized, but the middle and rear must be bored-to-fit in the block. This is a very important step, as the specifications for the cam bearing clearance are 0.001-0.0015 inch (0.025-0.037mm). The first task in optimizing the valve train, is to check the straightness of the camshaft. Knowing the bearing tolerance, the camshaft cannot be out-of-true by more than 0.0005 inch (0.012mm) or it will bind in the center bearing or have insufficient clearance. Next, the camshaft lobes should be in good condition and/or polished lobe surfaces . Ultimate friction reduction would involve DLC coating the bearing races and the lobes, to where the Ra value of the surface approaches 0.1.

Lifters and Lifter Bores - Next we need to look at the lifters and lifter bores. The foot of the lifter should have a slight convexity. This means a new lifter will have a slight crown to the foot. If you want to visualize this, take two new lifters and place the feet against each other. You will notice the curvature more readily when you do this. This curvature insured that the lifter turns a small amount each time the lobe lifts the lifter, to distribute the wear. If you ever find a lifter with a distinct wear pattern, then it likely that it has lost its convexity and is not rotating. In the race engines that I build in generally hone the lifter bores with a special small cylinder hone that has cork contact areas. I use a special compound, deposited on the cork and use this to provide a very light surface polish to the lifter bores. Again, if the ultimate in friction reduction is desired, then you could DLC coat the stem of the lifter. I do not recommend drilling any holes in the lifter. This significantly weakens the lifter and failure of the lifter can be the result. Be sure to inspect the push rod seat in the bottom of the lifter. This should have a shiny appearance and should not have any defects.

The Push Rod - The next item to consider is the push rod. I know that all of you will have seen advertisements for all types of alternative push rods, ranging from one aluminum ones to carbon fiber to metal-matrix-composite. Just keep the following in mind. The push rod should be as stiff as possible, in order to maintain proper cam lobe to lifter tracking, at a price you can afford. That said, yes, a push rod made from 3M aluminum matrix composite (not aluminum tubing) is lighter and stiffer than an equivalent 7mm 4130 steel push rod. However for most competitors it looks less appealing when the price of over $200 per push rod is realized. Because of the nature of a tube, as compared to a solid rod, a tubular steel push rod provides a good balance between performance, durability and cost. The last thing we want is for the push rod to flex. This produces unwanted harmomics in the valve train (valve spring to be exact) and will cause valve train failure.

If we divide the valve train into the components that are on the lifter side of the rocker arm fulcrum and those that are on the valve side of the rocker arm fulcrum, then it is more important to reduce the weight of all of the components on the valve side.

The Rocker Arm - Next is the rocker arm itself. Its weight is divided on either side of the fulcrum, and for most instances it is a bout a 50/50 proposition. Taking weight out of the tip side (where it contacts the valve stem) will help. Be careful how far you go, as you do not want to lighten it so much that the rocker becomes unreliable. In the rocker arm alone are three distinct boundary layer friction interfaces. There is the adjuster-to-push rod cup, the rocker-to-shaft, and the tip-to-valve stem.

To recap, we had a cam-lifter friction interface, a lifter-bore interface, a lifter cup-push rod interface, plus the three associated with the rocker arm. Not hard to see why this is such an important area, as none of these interfaces are hydrodynamic film interfaces* and have some sliding friction action associated with them.

The push rod cup-adjuster interface is one where little improvement can be made. I have considered making a very small orifice in the rocker arm which would exit on the push rod side of the rocker arm, behind the adjuster. This would allow some oil to squirt onto the adjuster and run down into the push rod cup. This would provide a level of oil cooling to this interface, as otherwise it simply replies on whatever random oil happens to splash into the cup. If you have a situation where the top of the push rod is discolored (brown or blue), then it is getting much too hot and additional cooling is required. This is generally a result of excess valve spring pressures and or lack of lubrication.

The actual rocker arm is the next item that we should take a look at. As you know, many used Fiat rocker arms have excessive play, even when installed on a new shaft. Even worse, as the metallurgy of the rocker and the shaft is similar, they both wear equally, making the problem more pronounced. I recently measured about 40 rockers and NONE had the required 0.002-0.0025 clearance, if used with a new shaft. New rockers are available (at about $40 per rocker) but are in limited supply. Of course a new rocker arm, without any metallurgical improvements, would suffer the same fate.

The problem of rocker longevity, and damage, is one that rarely affects standard road going Fiats, as these are generally not subjected to the extra stresses imposed by high lift camshafts and the associated parts. Yes, a street engine, poorly maintained with 50,000 miles on the odometer will have worn rocker arms for sure. In racing applications, where we run with stronger valve springs, much more aggressive lobe designs and much higher valve lifts, the dynamics are totally different. ALL of these put additional stress on the rocker/shaft boundary layer interface. As I mentioned before, even standard rockers from road cars will show signs of damage if the oil to the shaft has been less than adequate for some reason.

If we look at the dynamics of the rocker-to-shaft interface, the first rule of thumb to remember is that oil will take the course of least resistance. Therefore, if there is an excess of clearance, the likelihood of high spill flow is very likely. If there is no pressure (only flow) then, when spring pressure is applied to the rocker and all of the clearance moves to the top of the rocker, the oil flow will go to where the clearance is . Thus the very area that needs the oil, THE BOTTOM OF THE ROCKER, will not get any.

The second issue with the standard shaft, in a racing environment, is that the supply hole is too small to allow sufficient flow to service the 8 rocker oil delivery holes. As such, there is a pressure drop at this small inlet orifice and this will restrict the flow of oil. If, as I described above, there is excess clearance and the flow to the rocker is reduced, then all of the available flow will escape to the top of the rockers (where the excess clearance is) of the rockers immediately adjacent to the rocker shaft supply inlet hole. In this case, when the oil is hot, it is quite possible to have NO OIL PRESSURE build-up at the rocker arm/shaft boundary interface. This will then mean that surface friction will go up dramatically and heat will not be carried away as there is no oil flow at the rocker/shaft boundary interface. The end result is a seized rocker arm. (Note: Please read the section in blue later on for more clarification)

The solution is therefore to be found in four modifications.

1. Increase the supply hole to the shaft (remember there are two, even though one is used at a time) to 0.150mm (3mm) diameter. This is equal to the total area of the rocker supply holes** (0.027 inch [0.7mm] each) in the shaft. In this manner the input and output flow capabilities of the shaft will be balanced and insure that adequate flow and pressure are available to the rocker/shaft boundary interface.

2. Insure that the rocker-to-shaft clearance is approx. 0.002 inch(0.05mm), but no larger than 0.003 inch (0.075mm). This will insure that the supply flow to the rocker is greater than the "spill flow", and allow pressure to build up at the rocker/shaft boundary interface. This will promote better lubrication and cooling of this interface. As a secondary issue, the spill flow must be under sufficient pressure to cause oil to "splash" into the push rod cup. This is the only method for both lubrication and cooling this vital pressure interface. The surface area of the adjuster, where it rides in the push rod cup, takes the full force of the lobe opening pressure and must be adequately lubricated and cooled.

3. Modify the shaft, by providing a partial groove from the rocker oil delivery hole to the bottom of the rocker shaft, and then scoring an "X" at this point to spread out the oil. If the correct rocker/shaft clearance is maintained, then this will provide a wide cushion of oil for the rocker to work against.

4. Finally, where rules allow, the engine preparer may opt to install a small oil spray bar inside the valve cover, fed from a small line external to the motor. This will help insure that both the push rods and the valve springs are adequately oil cooled by high pressure oil. This is a common modification in many OHV motors, particularly in NASCAR, where flat tappet designs with very high valve lift are commonplace.

Further Information and Calculations

When a rocker is under tension, it is up against the bottom of the shaft and, there will be no oil flowing into the rocker/shaft boundary interface pressure point (presuming that the feed hole is on the bottom of the shaft). Considering that each rocker is under tension approx. 250 degrees out of every 720 degrees of crankshaft rotation, it can be assumed that, for the remaining 470 degrees of engine rotation, oil will flow through the rocker/shaft boundary interface clearance (approx 0.002 inch). This flow would re-establish a lubrication supply to the rocker/shaft boundary interface in preparation for the next 720 degree engine rotation cycle for that particular rocker arm, and, most importantly, carry away heat generated during the previous cycle at this boundary interface.

According to my observations there are 4 rocker arms in tension (to some extent anyway), at the rocker/shaft boundary interface, at any given point in time. During this "tension period" all the clearance is at the top of the shaft, with virtually nil clearance at the boundary layer pressure point (also the oil feed hole) on the bottom of the shaft. As 4 of the eight rocker feed holes are occluded, either partially or completely, any oil flow would be diverted to the remaining four rockers not under tension, with equidistant circumferential clearance.

Based on my earlier computations of cross-sectional flow area, each rocker's 0.070 inch (1.75mm) feed hole will flow about 10% greater volume than the spill volume of the 0.002 clearance between the rocker and shaft. The spill volume computes to an area of .004 sq. inch (2.58 Sq mm) per rocker arm. As such, the aggregate spill volume for 4 rocker arms would be 0.016 sq. inch. The single oil supply hole to the shaft must be 0.145-0.150 inch in diameter ,to supply sufficient flow, to service any four non-occluded rocker arms at any given point in time.

The metallurgical inconsistencies associated with a steel rocker arm against a steel shaft are not acceptable to for a high performance application. A better alternative would be a harder rocker shaft, made of 4130 steel and hard chrome finished for a surface hardness of RC65, and then an associated rocker arm with a pressed in bronze bushing. This combination of differing metals will prevent rocker arm seizures due to microwelding.

Scuderia Topolino has already undertaken steps to produce this new type of rocker shaft and rocker design.

With a conventional Fiat rocker arm the pad on the end of the rocker arm is larger than the stem of the valve that it contacts. Many times I have see this pad marked with a small “half moon” as indication that there is a great deal of pressure concentrated in a small area in this interface. I will talk some more about the causes for this pad damage in the camshaft dynamics later on. It has always been my contention that it would be better if we could make maximum use of the total surface area of the contact pad, and hence many of the motors that leave Scuderia Topolino have “lash pads or lash caps” over the end of the valve stems. This serves two purposed, both to protect the end of the valve stem and the pad on the rocker arm. Yes, it does add a minute amount of weight to the valve train on the “sensitive” side of the rocker arm fulcrum, but the benefits far outweigh the weight penalty. The pad and the associated lash cap should be polished to a fine finish to reduce parasitic losses.

One other possibility is to use an alternative rocker arm. Scuderia Topolino has available an aluminum rocker arm with a rollerized tip. The aluminum rocker arm body is made from 2024-T6 aluminum and is then hard anodized. This provides a surface hardness of RC62, and combined with the ductility of 2024 aluminum, provides a usefull wear surface against a steel shaft.

* There is some evidence that at higher RPM the lifter-cam lobe interface converts from a boundary layer interface to a hydrodynamic interface.

Valve Spring Retainer – Here is the first area where we can make a real weight savings. By changing to a titanium retainer we can cut the weight in half. It may not seem like a great deal, but it will make a decided difference in the overall dynamics. Scuderia Topolino uses special retainers that use special collets, with an included angle of 6 degrees, as opposed to the standed Fiat ones which are 5 degrees. We can also provide these retainers with 7 degrees collets, and then the collets can be provided in titanium as well.

Valve Springs - We finally pay attention to these when we install a new camshaft, to check that we do not have coil bind. There are however a number of other considerations that must be examined in terms of valve springs.

Most standard valve springs are selected on the basis of basic performance and longevity. Obviously when the 4 cylinder Fiat engine was first developed, what with a whopping 27 horsepower, the requirements of valve spring performance was not very comprehensive. After all, with a camshaft with a total lift, at the valve, of less than .300 thousands of an inch (7.6mm), the real criteria was to use a spring that would last a long time. These springs had a sufficient number of coils so as to not be very highly stressed.

In a racing situation the requirements are just the opposite of that of the standard road car. First, whereas the standard engine uses relatively low RPMs, therefore spring harmonics play a minor role. Not so in a competition engine. Modern racing camshafts not only open the valve more and for a longer period, they also open the valve quicker as well. This means that the modern valve spring must be able to control rapid valve train acceleration. We need to find a good balance between spring pressure, harmonic control and minimum parasitic loss. Without a great deal of trial and error, the only other alternatives is to computer model the valve train, or to follow the recommendation of the camshaft supplier. The cam grinder will almost ALWAYS be very conservative in choosing a valve spring. In most cases he would rather err on the side of a too heavy spring, than one that is too light.

Valves – Here again, we are confronted with a choice between ultimate weight savings and cost. One way to reduce the weight of the valves is to use a small diameter stem, say from 7mm to 6mm. Also, many racing oriented valves, because the way they are formed, are inherently lighter than the standard Fiat valves. Finally, if the pocketbook will allow, you could go for titanium valves. This will greatly reduce the weight. I am fairly comfortable with using titanium intake valves, but titanium exhaust valves have a very short lifespan, and should be replaced at least once each season. This of course adds to the cost.

Valve Seats – While technically there is nothing that can be done to lighten valve train components with the valve seat, if you plan on using titanium valves, then there will be a cost impact that is not insubstantial. Unfortunately you cannot use steel valve seats, as the titanium valves are subject to a phenomenon call “microwelding”. Tiny amounts of steel are transferred to the valve during operation and eventually the valve no longer seals. This means that either special copper or berrylium valve seats must be used. Unfortunately these are from 6-15 times more expensive than a steel seat.

4.2 Camshaft Dynamics

Camshaft and Induction Dynamics

In an effort to better understand camshaft technology, I asked a number of fellow Abarth competitors what type of camshaft they were using. I wanted to take these various camshaft designs and put them through a computer based engine simulator to compare the various grinds. Below you will find a spreadsheet of the information that was provided.

Manuf

Adv Duration

Lobe Center

Cam lift (mm)

Intake L/C

Overlap

Open/Close Adv.

CatCams

305

108

7.21

108

78

39/79 79/39

PBS A8

305

108

6.98

108



PBS A6

292

106

6.85

104

72

40/72 80/32

SLR286

286

106

8.43

106

74

37/69 69/37

SLR300-106

300

106

8.43

104

88

46/74 78/42

SLR300-110

300

110

8.43

105

80

45/75 85/35

Kent FT6

304

106

7.11

106

92

46/78 78/46

Alquati

316

110

6.85

108

96

48/88 88/48

Abarth

316/304

105

7.21

102

100

53/83 77/47

Abarth

336

105

7.72

105

120

60/96 96/60

 

As you can see, almost every competitor is using a different camshaft. I did find that two people were using the same Kent FT6 camshaft. Duration ranged every where from 286 to 336 degrees. Certainly there were some interesting surprises when I ran some of these profiles through my analysis program. My first reaction was that the larger the duration, the more horsepower. Not quite correct. At the end of this study I will rate each of the above cams that I was given "advertised" duration information for.

First I have to state the assumptions that I used to make all of the comparisons that follow.

Bore

68m

Stroke

74mm

Rod length

110mm

Compression

13.8:1

Cylinder head

PBS 8P

Inlet Valve

31mm

Exhaust Valve

27mm

Carburetion

2 x DCOE40

There is a complex "ballet" of inter-acting numbers that define a particular camshaft and what it is capable of delivering. As with anything in life, a good "plan" is always worth the time it took to develop. Such a plan is also required when deciding upon a camshaft and the other components it will be required to interact with. In my mind there are three principal elements. It is very much like a 3-legged stool. It cannot stand unless all three parameters are well thought out and developed.

A) The maximum RPM that will be used. - It is of little value to specify a camshaft where horsepower is likely to be produced above the RPM range where it will be required.

B) The RPM where maximum torque will be expected. - This may have more to do with shift points and such, but it is something that can be adjusted for to some extent.

C) Best Available Octane Fuel – This will affect which camshaft can actually be used without encountering detonation.

From the chart above, it is obvious that different competitors had different things in mind when choosing a camshaft. I am of the firm belief that, given the design of the small Fiat motors with only three main bearings, that a top RPM around 8500 is not only prudent, but the only way of assuring reasonable reliability. Therefore, I will go out on a limb and say that anything over 304 degrees duration is may too great, as it puts the power production too high in the RPM range. Now there will be people who disagree, but don't dismiss the idea just yet.

The PBS head and intake manifold/Weber carburetor combination has an overall inlet tract length of around 330mm. Generally the carburetors would have 30mm chokes. We want to maintain a mean a peak torque number somewhere between 5500-5800 RPM. Below is a chart of the calculations that I did for a 1050cc motor (68mm bore 74mm stroke, volumetric efficiency of 85%) for the diameter of the inlet port at the head/intake manifold interface.

    

Peak torque RPM     Inlet Diam.(mm)
5000     21.8
5250    22.8
5500    24.1
5750   25.1
6000  26.2
6500  28.4
7000  30.7

All PBS heads are machined for a 25.4mm port, so it would appear that there is a good match. This size port assumes an air velocity of just under .6 mach, or 650 ft/sec (196m/sec). So to match these head characteristics we will be looking for a camshaft that will deliver peak torque between 5000 and 6000 RPM.

Finally we need to look at third leg of our three legged stool, namely Fuel Octane Rating, and this will then tie directly to Dynamic Compression Ratio (DCR), which is quite different from the Computed Compression Ratio (CCR). We should all be familiar with how CCR is calculated. Basically it is the displacement of the cylinder, plus the displacement of the combustion chamber, then divided by the volume of the combustion chamber.

For our "standard" engine I stated that the compression ratio would be 13.8:1. Of course the intake valve is NOT closed for the entire compression stroke (BDC to TDC). In fact, even though the piston has already passed BDC, air is still flowing into the cylinder due to scavenging effect. The key is adjusting the intake valve closing position so that it coincides with the point where intake flow ceases.

According to the camshaft specifications in the chart the intake closing is anything from 65 to 96 degrees. With such a wide variance, it will be interesting to compare.

A small story - At a recent race at Hockenheim I ran across Bram Paardenkoper. During our conversation he mentioned that he had installed an Alquati 316 degree cam in his car for the weekend, as he was trying to find some extra pace. He indicated that over the years he had done quite well, but now the competition was starting to reel him in. He was however disappointed as the car, with a cam with 12 degrees more duration, was actually slower than before!! The Alquati cam is very similar to the 316 can in the list, and the cam Bram used previously is very close to the Kent FT6 grind. This led me to do some further analysis.

I started out by looking at where the rod and crank throw were at 90 degrees to each other on the compression stroke, in terms of degrees of crankshaft rotation, and how many degrees this was from the point of maximum piston acceleration. Perhaps this diagram will help in understanding the the reason why the rate of piston acceleration is not uniform over the entire stroke.

It turns out that for the A112 engine this is at 71.4 degrees before TDC. Conversely, this means that on the compression stroke the point of maximum acceleration for the piston is 108.6 degrees from BDC.

The most important consideration for any engine is the timing of the closing of the intake valve. The valve must be closed before this point of maximum piston acceleration to minimize the effects of pressure reversal in the cylinder. Of course the earlier that we can close the valve, and still meet our design objective with regard to power generation, the better the performance will be. After modeling many camshafts for use in the A112 motor, with its particular bore and stroke characteristics, it would appear that a valve closing period between 65 and 78 degrees ABDC produces the best results, at least for engines like the A112. Coincidentally this also places the closing event prior to the crankpin achieving a 90 degree angle with the connecting rod centerline, so that the intake valve is closed prior to the fastest portion of the piston acceleration.

The challenge is then to find a combination of lobe center, overlap and duration that will maximize DCR. Remember that DCR can only be the amount of the stroke from the time the intake valve is closed to TDC.

Suppose we have a cylinder volume of 268.75cc per cylinder (68mm stroke x 74mm bore), with a combustion chamber volume of 21cc.  This would provide a  CCR of 13.8:1.  Now if we use the Alquati 316 degree cam (110 deg L/C, 6.86mm lift at the cam), with an intake closing of 83 ABDC, we can compute the position of the piston above BDC and determine the actual cylinder volume remaining at that point.  In this case, at 83 degrees the cylinder volume (VE) is 170cc.  So according to the formula "(Cylinder Volume + Combustion Chamber Volume)/Combustion Chamber Volume", this makes for an  DCR of 9.2:1.  The DCR is always lower than the CCR, it is only important how much lower.

Now if we take the Kent FT6 304 degree camshaft (106 deg. L/C, 7.11mm lift at the cam) with a intake closing of 74 degrees (installed 4 degrees advanced) this results in a VE of 192cc and a DCR of 10.2:1. This is a full compression point higher. In addition the overlap on the Kent camshaft is also 4 degrees less.

Small story continued - Seeing the above results I could easily see that the Kent camshaft was a much better cam for the A112 motor.

Most camshafts are installed 2-4 degrees advanced, particularly if they use a cam chain which over time will stretch. This will change the Effective Cylinder Volume and also the Effective Compression Ratio. Below find a spreadsheet of the results of the computations for the various camshafts. Remember that Actual Cylinder Volume is 268.75cc per cylinder.

Manuf

Adv Duration

Lobe Center

Cam lift (mm)

Intake L/C

Overlap

Open/Close Adv.

VE in cc

CRE

CatCams

305

108

7.21

108

78

39/79 79/39

182.4

9.68

PBS A8

305

108

6.98

108





PBS A6

292

106

6.85

104

72

40/72 80/32

196.88

10.37

SLR286

286

106

8.43

106

74

37/69 69/37

202.73

10.65

SLR300-106

300

106

8.43

104

88

46/74 78/42

192.86

10.18

SLR300-110

300

110

8.43

105

80

45//75 85/35

190.82

10.09

Kent FT6

304

106

7.11

102

92

50/74 82/42

192.86

10.18

Alquati

316

110

6.85

108

96

48/88 88/48

162.31

8.73

Abarth

316/304

105

7.21

102

100

53/83 77/47

173.67

9,27

Abarth

336

105

7.72

105

120

60/96 96/60

143.34

7.83

By looking at the above chart it becomes obvious that BIG duration, with small lobe centers, equates to a severe lowering of VE and the directly interrelated DCR. So a compromise must be reached. We can choose a larger lobe center, making sure the power band stays where we want, and also advance the cam to get back the VE and DCR numbers that we want.

This is what has been done with the SLR cams, Kent FT6 and the PBS cam and it appears that the "sweet spot" is between 71.4 and 78 degrees for intake valve closing. This produces a DCR above 10 in all cases. Noting is free however, and DCR numbers over 10:1 may give cause for alarm, as you would definitively have potential for detonation. It may be that slightly lower CCR may have to used, so as to lower the knock index number.

Note: I went back to an article that I had read some years ago about the short lived Pontiac GTO Trams-Am project. Pontiac's project engineer Tom Neil explained how they had gone about determining what their "road-race" engine required in the way of a camshaft. At the end of the day they also determined that the "secret" lay in the closing times of the first the intake valve and secondarily the exhaust valve. As it turned out they settled for a 300/310 camgrind on 105 centers with ,500 inch (12.5mm) lift. Intake closing was slightly different, because of the short stroke and long connecting rod. However when computed backward, it falls right into the range that I found to be effective for the A112 motor.

There is one additional difference and that is that the SLR cams all have smaller exhaust/intake overlap and appreciably greater valve lift than the other cams in the list. This means the lobes on these cams will be more aggressive as far as lift per degree of rotation and so exhibit greater lift "earlier", providing a greater area "under the curve". I used one of my engine design packages to illustrate this in graphical form, as in number form it becomes too cumbersome.

First, let's compare the two most directly comparable camshafts on the list for which I have data. This would be the Kent FT6 and the SLR300-106. These are 304 and 300 degrees respectively (both on 106 L/C), with the Kent FT6 winning out in duration and the SLR on valve lift and less overlap. Lets see how they compare.

Interesting !! Here you can see that cams with the same VE and DCR, can have different power curves. Obviously the higher, and inherently earlier, cam lift has increased the area under the curve for the SLR300-106 cam. The SLR300-106 cam also has 4 degrees less overlap. Both the horsepower and torque show an reasonable increase, about 5 HP and 2 lb/ft of torque.

What would happen if we spread the lobes apart to 110 degrees, thereby reducing the overlap by another 8 degrees? Would this reduction in overlap, theoretically reducing pumping losses, have beneficial effects?

From the results you would have to say that the SLR300-110 cam does appear to do better. While the horsepower increase is relatively small (maybe 1-2), it has moved the horsepower peak higher in RPM, and there is also a noticable increase in peak torque, even though the VE and DCR numbers are marginally less than the SLR300-106.

What about the SLR 286 cam (also known as the Max II). It had higher VE and DCR numbers than the other two SLR300 cams and the Kent FT6, yet it is another 4 degrees less in duration. How does it compare?

Did you expect this result? Let's see if we can judge why. The torque is considerably higher, based on having the lowest overlap of any of the cams we tested. Likewise, because of its shorter duration, it does achieve peak horsepower as early as 6500 RPM, and by 8500 has dropped off considerably.

As far as the original design criteria of a cam that had peak torque between 5000 and 6000 and carried power to between 8000-8500, this may not be the best cam, although it is what I used in 2001 when I ran in the Coppa MIlle. On the other hand if you were doing fast slaloms or perhaps even hill climbs, where good torque response at lower RPMs are required, then the SLR286 would get my vote.

Small Story Final Episode - Just to finish the story, let compare the Kent FT6 cam that Bram would regularly run with the 316 degree Alquati camshaft that he decided to try at Hockenheim.

Well as you can see the Alquati cam comes up well short of the original Kent FT6 and all of the SLR designs would probably outperform the Alquati as well.

So to summarize, it is probably more important to maximize the "area under the cam curve", by increasing both the aggressiveness and lift of the cam lobe and making sure that the valve closing is somewhere in the range of 65-78 degrees ABDC. Likewise, using slightly larger lobe seperation (108-110 degrees) will also reduce overlap and minimize possible pumping losses.

4.3 Optimizing Valve Train Reliability

  1. Valve Spring Forces - The answer is both simple and complex. The simple one is "enough to keep the valve from floating or bouncing off the seat". The more complex answer takes into account the weight of various components and the aggressiveness of the opening and closing ramps.

    In order to answer this fully you would have to run the engine on a Spintron machine. Then with a high speed camera and a strobe you could isolate each of the movements of the camshaft action and the impact on the valve and spring. Since almost none of us have access to this type of equipment, we have to make some educated assumptions. These are that we need sufficient spring pressure to make sure that the lifter accurately follows the cam lobe and that the valve is not lofted off the lobe. (There are some cam designs where the valve is PURPOSELY lofted off the valve lobe in order to achieve higher opening lifts, however this may have other consequential effects which under normal circumstances would be catastrophic.)

    We already know that with medium lift camshafts (6-7mm at the lobe) that standard valve springs are quite adequate. Abarth made some springs that dealt with cam lobe lifts of 7-7.4mm. Finally, with aggressive camshafts with 7.5-8.2mm lift we need springs with both more preload, and with more free travel. The more aggressive cams will have valve openings between 11.5 and 12.3mm.

    I have seen some valve installations where the seat pressure was less than 30 lbs and the nose pressure only 120 lbs. The Abarth springs are rated at about 45 lbs seat pressure with about 160 lbs across the nose. For very aggressive camshafts seat pressures go up to about 60-70 lbs and nose pressures may be in excess of 230 lbs. This clearly illustrates that the rocker/rocker arm lubrication boundary interface is being stressed much harder with aggressive camshafts and springs.

    We can do things to minimize the spring pressure however. First and foremost would be to reduce the weight of those items on the valve side of the rocker arm. This includes the valve, retainer, spring and the rocker arm itself. In a secondary fashion the valve spring is also responsible for controlling the lifter contact with the camshaft. It is in this area where there may be significant gains that can be made.

    If we were to make a spring holder that would sit on top of the lifter, with an orifice for the push rod to go through, we could put a compression spring there to control the action of the lifter, and a portion of the push rod. The other end of the lifter spring would be held captive by a notched plate on the underside of the cylinder head, and positioned in the lifter galley. In "hot-rod" circles this would be referred to as a "rev-kit". This would relieve valve spring of this responsibility and the spring tension of the valve spring could be reduced. This would also mean reduced pressure on the adjuster/pushrod and rocker/rocker shaft interfaces, thus reducing the parasitic losses associated with these interfaces.

    The resultant redistribution of spring pressures would either allow more RPMs before valve float would occur, or would allow lighter valve springs to be used with the current limit on RPMs being observed.

    My current test have indicated that if a spring is mounted on the lifter, then the valve spring pressure could be reduced by as much as 25% or more.

    As with most things in life, there are really no "free lunches", and the same goes for this idea. It is true that the redistribution of spring pressures would have a beneficial effect on the rocker arm/shaft boundary interface. However there are some trade-offs.

    1. 1. The spring in the lifter idea will add three extra components. All of these items will add to the weight of the total valve train.
    2. 2. The addition of a third spring will also add a third harmonic element to the valve train. So we have a dampened pair of springs on the valve and an undampened spring on the lifter.
    3. 3. The chilled iron lifter will now be constantly spring loaded on the cam lobe. Principally this will have an effect on the heel, or base circle, element of the lobe. This means that additional lubrication may be required to deal with this. As the standard lifter tends to collect oil, it would be possible EDM a small hole in the bottom of the lifter to provide a "drip" oiling system for the cam lobe. It is not known if the metallurgical structure of the chilled iron lifter would take well to any "interruption" in the lifter foot surface, no matter how small. It could be that this "defect" could cause be the beginning of catastrophic lifter failure, and also perhaps camshaft failure.

    Having looked at all of the "pros and cons", I have come to the conclusion that this is not an avenue that is worthwhile pursuing, at least at this time, simply because the attendant risks outweigh the possible rewards.

    However, what it did reinforce was that a better understanding of the spring forces required for the proper operation of a camshaft is required. To that end I went back to one of my engineering programs that allows me to model for valve train dynamics. This involves recording the lift of the camshaft at 1 degree intervals (Ideally you would want to record the data at much smaller increments with a Cam Doctor or similar device, but for this exercise this level of accuracy will be sufficient) and recording this information in a "camshaft file". This information, along with the weight of each individual component in the valve train (valve, spring, retainer, collets, lash cap, rocker arm, push rod and lifter) can then be modeled to determine what the minimum seat pressure and full lift pressure that are required to have accurate tracking of the lifter to the cam lobe.

    At this point I have to again indicate that all of this came about because of a lubrication problem at the rocker arm/shaft boundary interface, but it also important in terms of the overall performance of the engine.

    For examination purposes I used our SLR300 camshaft as our test sample, principally because this is the camshaft where the rocker arm/shaft lubrication problem first emerged and it would appear to be linked to the increased spring pressures exerted by the uprated springs, supplied by Scuderia Topolino, for use with the SLR 300 camshaft. Remember, the primary valve spring criteria is that the cam must not float the valves within the operating range of the engine.

    So just how much spring pressure is required to control the valve action with a SRL300 camshaft, yet NOT overstress the rocker arm/shaft boundary interface? I decided to draw on some empirical data, provided by customers, and to model all of this to try an determine an answer. To this end II had to set some standard weights for the various valve system components. These are within one gram +/-.

    Push Rod 48gm, Lifter 29gm, Intake valve 44 gm, Exhaust valve 42gm, Titanium retainer and collets 8 gm, Spring 50gm, Lash cap 3gm, Rocker arm 52 gm.

    One of my customers (Customer A) used the SRL300 camshaft with a set of the valve springs we supplied, rated at 90 lb (41 Kg) seat pressure and 245 lb (111 Kg) full lift pressure. This engine suffered from a rocker arm/shaft lubrication problem, with the far rear rocker seizing to the shaft. Without knowing the condition of the rocker and shaft in terms of wear and clearance when the engine was assembled, it is difficult to estimate just how much the limits had been exceeded. This is of course always a problem with high performance camshafts, as the cam designer has no knowledge of the condition of the remainder of the components that must work together to make an effective camshaft installation. Obviously, this amount of pressure, given the Fiat lubrication system, was too much and the rocker arm failed.

    Another customer (Customer B) used the SRL300 camshaft with a set of Schrick springs rated at 50 lb (22.7 Kg) seat pressure and 151 lb (68.6 Kg) full lift pressure. This engine did not suffer from a rocker arm/shaft lubrication problem. From this empirical result it would be safe to assume that if the rocker arm/shaft boundary interface were to factory clearance specifications, that a rocker arm failure would not occur. *

    Obviously, the answer must lie somewhere in between.

    I modeled both the intake and exhaust valve valve lift and actual dynamic valve lift characteristics for Customer A first. This was a SLR300 camshaft with the Scuderia Topolino supplied valve springs. Test RPM was 7500 RPM.

    As you can see in both graphs the two traces for (red and green) are perfectly superimposed. This would seem to indicate that the valve spring supplied by Scuderia Topolinbo has sufficient pressure to control valve. What it does not say is whether the spring pressure could be less, and still do an adequate job of controlling the valve train motion.

    Next I modeled the same SLR300 camshaft, as used by Customer B, using the spring specifications for the Schrick valve spring that he used, again at the same test RPM of 7500 RPM.

     

    Here the result is quite the opposite. The two traces are easily distinguished. Both the intake and exhaust valve are being "lofted", indicating that the lifter is not properly tracking the cam lobe. This is not necessarily "valve float" in the traditional sense when the spring goes into a harmonic condition, but it may in turn cause some unwanted spring oscillation to occur. Obviously this is spring is not sufficient to control the valve train motion of the SLR300 camshaft. The lifter is for all intents off the closing side of the lobe altogether and comes back down with the lobe at less than 0.050 lift. You can also see that it is predicted that the valve will bounce when it does come down with considerable destructive force.

    * The engine was assembled by a knowledgeable mechanic. The cam was degreed in and found to be installed to specification, with minimum valve-to-piston clearance of 0.100 inch (2.5mm). He did set the valve clearance at 0.008/0.010 inch (0.2-0.25 mm) instead of the recommended 0.020/0.022 inch (0.5-0.55mm), as he thought the clearance was too great and "could not be right". The closer clearance of course has two effects. 1) It effectivecly increases the duration of the camshaft about 20 degrees. 2) It compromises the effectiveness of the "lobe clearance take up ramps". 3) This combined with the lower pressure Schrick valve springs led to the lifter crashing on the end of the closing ramp and bouncing off the seat.

    This engine did lose compression in 3 of 4 cylinders after about 15 minutes of running. What can we deduce from the above information? First it is obvious that both the intake and exhaust valve are not in contact with the cam lobe on the closing ramp, and totally overshooting the clearance take-up ramp on the camshaft. This would cause a very high spike in pushrod pressures. (Note: On engine disassembly the camshaft had at least one damaged cam lobe) See the following graphs.

    These graphs illustrate the effect of valve "lofting". The closing ramp forces are more than three times higher than the opening ramp forces. This camshaft was going to fail, it was only a matter of time.

    Valve lofting has other unintended consequences. First, the inertia contained in valve lofting could cause the associated valve spring to go into coil bind, because it does not have sufficient tension. This will place severe stresses on the retainer and collets. Further, if the amount of lofting is in excess of the assembled piston-to-valve clearance, then almost certainly the exhaust valve will come into contact with the piston. From the spring technical data it appears that the spring can compress an additional 3mm before it is in coil bind. If the valve is lofted into spring bind (.120 inch [3mm] more than the actual lift of the cam lobe), then there would be interference of about 0.020 inch [0.5mm]) between the piston and valve. This situation is more acute for exhaust valve, due to the motion of the valve relative to the piston and would be worse if the cam were installed advanced 3-4 degrees to accommodate any chain stretch. The test data suggests that the valve lofting will occur between 7500 - 8000 RPM if the Schrick spring is used with the SLR300 camshaft.

    Conclusion - If the engine was assembled with a valve-to-piston clearance of 0.100 inch (2.5mm), then certainly at 8000 RPM their is a very good likelihood that the pistons will clash with the exhaust valves. The problems associated with the engine of Customer B seem to bear this out.

    Note: Under no circumstances would I imply that there is physically anything wrong with the Schrick valve spring. This company makes very good products. It simply means that it is not the correct valve spring for a SLR300 camshaft. In fact, in other tests that I conducted with different camshafts with less duration, there appeared not to be a problem at all with this spring. The margin of difference is very small.

    So what is the answer.

    After further modeling, if a valve spring with a rate of 220 lb/inch (3.9 Kg/mm) were used, then lofting is not longer an issue. Please note that there is a difference between the “spring rate” and the actual seat spring pressure and the over the nose pressure. As a matter of safety, I would probably opt for a spring with a rate between 250-260 lb/inch (4.47-4.65 Kg/mm) to provide a little extra safety margin. This valve spring would have 52 lbs (23.6 Kg) of seat pressure (at 1.250inch [34.29mm] installed height) and 172 lbs (78 Kg) of pressure at 12.18mm valve lift.

    This spring combination would be a 24% reduction in valve spring force, as compared to the spring supplied by Scuderia Topolino, and recommended by our cam grinder. This is not the complete story however. As the rocker arm ratio is 1.45:1 the reduction in pressure on the valve adjuster is close to 35%.

    It goes with out saying that a reduction in spring pressure, as well as the attendant reduction in other parasitic friction losses associated with using a spring with less pressure, will mean an incremental increase in horsepower. Between 7500 and 8000 RPM this may very well mean 2-3 more horsepower is available to drive the wheels.

  2. Rocker Arm Geometry – One of the misunderstood and mostly forgotten functions of the push rod is to adjust rocker arm geometry. After the cam has been reground, the head and block decked, new valves installed in the head, why would you assume the standard push rods are still the correct length? After all regrinding the cam lowers the base circle of the cam, thus lowering the position of the push rod to the rocker arm. Now, if you also machined the head 0.040 inch (1mm), and perhaps the block also 0.020 (0.5mm), then you will have compensated for some of the material ground from the base circle of the cam. The point is, you never know how much of each, so when the engine is first test assembled is when you find this out.

    As a good rule of thumb, when the cam is at 33% of its total lift (so for the SRL300 that would be 0.110 inch [2.8mm]) the adjuster should be in a straight line with the pushrod, with the adjuster showing no more than three threads below the rocker arm. So if 5 threads are showing, and each thread is .8mm, then you would require a push rod that is approx. 1.6mm longer than the test push rod. If however the cup of the push rod is right up against the rocker arm, then you will need a push rod that is 2.5mm shorter than the one you are using for the test.

    Note: The easiest way to measure the effective length of a push rod is to put a small ball bearing in the cup and then measure the overall length with the ball bearing. Now measure the diameter of the ball bearing and subtract this amount to get the effective length.

    You are not finished however. Next you must look at the rocker arm pad where it contacts the valve stem. Rocker arm geometry is generally optimal when the travel or movement of the rocker arm tip on the valve stem is minimized. To understand how to achieve correct geometry, it must be understood that the rocker arm tip itself travels in an arc. At zero lift, the rocker arm tip is expected to be closer (or inboard) to the plane of the pivot point and as the valve starts moving down, the rocker arm tip starts moving outboard. If the geometry is close to ideal, then the rocker tip will be at its most outboard position at half or mid lift at which point the rocker tip starts moving inboard again as the valve reaches full lift. Simply put, ideal rocker arm geometry is achieved when the rocker tip is sitting on the valve stem tip at the same position at both zero lift and full lift.

    In a perfect world, where the rocker shaft pedestal stand locations, the valve guide, and the rocker itself are all machined to exact specifications, the rocker tip is expected to be sitting slightly inboard of the valve stem center at both zero and full lift while the rocker tip will be sitting the same distance outboard of the center of the valve stem at exactly mid-lift. As this is not a perfect world, this sometimes does not happen.

    It may be that the valve is installed at a slightly lower installed height. If so a lash cap may be required. Alternatively, if the rocker pad does not sit in the correct location on the valve stem and the rocker fulcrum point has to be moved closer to the center of the stem, and so a shim will have to go under the rocker arms stands. If the opposite is true then the rocker stands may have to be reduced in height a small amount.

    All of these actions, will affect the length of the push rod required. Take the trouble to do it, as it is worth the effort to get the geometry correct.

    Note: The condition of having the contact pad too far extended over the valve stem, is the cause for the little half moon wear marks, that I discussed earlier. This causes the rocker arm to extend “over” the valve stem at full lift, and instead of depressing the valve stem it pulls the valve stem sideways toward the rocker arm fulcrum. This increases parasitic friction losses and causes premature valve guide wear.

 
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