We have finally arrived at the socalled cork in the system, as far as power production is concerned. As Chart 6-1 shows, optimizing airflow (and that means getting as much as possible) is the number-one goal. But as far as heads are concerned, it is not the only factor in making big power numbers. To understand the technicalities involved, we start with the need for good airflow. Without this, it matters little what other factors are imperfect; the engine makes less power, no matter what.
This Tech Tip is From the Full Book, DAVID VIZARD’S HOW TO BUILD HORSEPOWER. For a comprehensive guide on this entire subject you can visit this link:
SHARE THIS ARTICLE: Please feel free to share this article on Facebook, in Forums, or with any Clubs you participate in. You can copy and paste this link to share: http://musclecardiy.com/performance/horsepower-secrets-cylinder-heads/
Optimizing Cylinder Head Airflow
With the classic poppet-valvestyle engine, air cannot make a straight shot into the cylinder. There has to be a bend in the port to accommodate the valve stem and, worst of all, the air must make its way around the head of a valve. It is the head of the valve and the seat on which it rests during the closed position that is the worst impediment to flow. Although I have used it many times before, Illustration 6-3 serves well to put into perspective the relative flow restriction each section of a typical production port has.
Within either port, the valve is the only component that, as it opens and closes, delivers a variable geometry. At low lift, the flow is entirely dependent on the size of the gap between the valve seat, the seat in the head, and the efficiency that it flows at. At low lifts, the geometry of the seat and the area closely surrounding it play a huge role in dictating its flow efficiency.
At high lift, the port size and shape dictate the amount of flow delivered. This being the case it’s helpful to have an idea of how the flow priority changes from seat form to port form as the valve goes through its lift cycle. With the aid of a flow bench and some calculations, this can be determined; the result (shown in Chart 6-4) is that the valve seat influences flow to a much higher lift than you may intuitively suppose.
In Chart 6-4 two lines cross (in the circle) and, at this point, the form of the seat is equally as important as the size and shape of the port. Although these results were from a production pushrod cylinder head, they can be quantified to apply reasonably well to any style of head— two- or four-valve; stock or race. This can be done by looking at the lift-to-valve diameter ratio. In other words, we are reducing things to proportions rather than absolute dimensions.
A key proportion is a lift value equal to one quarter of the valve’s diameter. This is commonly known as 0.25D; Illustration 6-5 shows the significance. At 0.25D, the “curtain area” is exactly equal to the valve head area. By applying the same criteria, we can say that the valve seat influence, though continually diminishing, is the major influence on flow, up to a valve lift equal to about 0.18 to 0.20D. In other words, the seat is the number-one priority until the valve has reached a lift figure between 18 and 20 percent of the valve’s diameter.
At this point, you may well ask just how much difference a valve seat shape can make to the efficiency delivered at the lower lift. The answer is: a lot. Up to the 0.20D lift point, a typical production three-angle valve seat usually delivers efficiency figures starting at about 65 percent and quickly dropping to around 55 percent. Spend time (a lot of it) on a flow bench, and those figures can get as high as 85 percent and sometimes more.
A point worth raising here is that there is a commonly held belief, among many successful engine builders who specialize in high-output two-valve V-8 engines, that too much low-lift flow hurts power. I won’t go into a lot of detail here other than to put forward a simple explanation. If you want a highly detailed analysis as to why this is not so, then go to my porting school series at www.motortecmagazine.com.
The simple explanation is: First, the cylinder does not know how far an intake valve is open. All it knows is how much flow is being presented. Therefore, if too much flow is being presented at any given moment, it is because the cam valve events are not what the head/engine combination wants. That means the wrong cam is being used. Claiming a cylinder head has too much low-lift flow is a little like claiming your race team has too much money!
In Illustration 6-6 “A” is a knife edge seat (the port in the head is the same size as the seat). This gives the maximum area for air to enter the cylinder at high lift, but the sharp corner is very un-streamlined. That means even if we could seal the valve in this instance, it still would not flow well.
The obvious answer to the sharp corner is to chamfer it with a 45-degree cut as in “B.” Even though the area directly under the valve is smaller, the combination flows better. If we work on the principle that some streamlining is good, then maybe more is better. However, we find that a second angle at 60 degrees beneath the 45, as in “C,” improves this even further. Finally, a small exit cut on top of the valve seat at 30 degrees further helps the situation, as shown in “D.” All this tells us that the seat and the port in the close vicinity are all about size and streamlining.
This brief description of the developed shape of the seat does tend to make things look easy, but thinking so is a mistake. As we shall learn later, flow around a valve is not even because of the direction of its approach. This means that the simple valve seat form we have just discussed is probably okay but hardly optimal. To get a near-optimal form with any given port requires a flow bench and usually a fair amount of testing.
You may have heard of valve shrouding; it and its effects are closely tied with the seat and port design that works best. For that reason, let’s look at it now rather than later.
So what is valve shrouding? It is easier to see from a drawing than to explain, so take a look at Illustration 6-7. This is an intake valve on the intake side of a combustion chamber. The point to note here is that the walls of the chamber are, for much of the valve’s circumference, close to the edge of the valve. This shrouds that part of the valve, limiting flow around that particular section. The green line represents the radius of the stock chamber wall, and the airflow produced by this is shown by the green line on the chart in 6-8. We obviously cannot cut the chamber away where it is adjacent to the bore diameter (gray line), but elsewhere we can cut it to alleviate, as far as possible, the shrouding caused.
We can look at the chamber wall as a continuation of the valve seat. At an angle of 36 degrees from the valve seat on up, the area around the valve is always as much as the curtain area. This represents a geometrical un-shrouded valve.
Illustration 6-9 shows a typical two-valve-per-cylinder, wedge-shaped combustion chamber. It is already limited on breathing capability simply because it is only a two-valveper-cylinder design. It makes no sense to further limit this type of head’s power-production capability with needless shrouding.
All this talk of shrouding raises the question as to whether it is possible to have zero shrouding within a head that still utilizes the largest valves possible. A hemi-style combustion chamber can provide just that. The reason it does so is because the valves are always moving away from the cylinder wall as they open.
Many World War II aero engines had a hemi-style combustion chamber. These typically employed valve angles of as much as 90 degrees inclusive. This accommodated the biggest valves, but it also produced a very deep chamber (half a hemisphere). Deep chambers were okay for the low-compression ratios used for heavily supercharged engines. But they were bad for high-compression use because of the high piston dome needed. In practice, it turns out that the optimum angle for the intake valve from the bore centerline is about 18 degrees. For the exhaust, where shrouding is less important, the optimum angle is about 10 degrees.
Chrysler has been synonymous with hemi engines from the 1950s and with good effect. Their latest iteration, introduced in 2003, is the 5.7 (and a 6.1 version that came later) engine. The engine is a very well-conceived design with heads that flow every bit as well as you would expect a good design to do.
So far, we have talked about geometric shrouding. It’s a good start to understanding what shrouding is. But simply applying it, without further thought as to other factors in the combustion chamber, is not the way to go. If air entered all around the valve in a uniform fashion, then geometrically un-shrouding the valve would work every time. However, air is heavy stuff and tends to want to flow in a straight line. There is no point in un-shrouding a portion of the valve’s circumference if there is minimal flow there, so we must first understand where the unshrouding needs to be done to make the most use of material removed from the combustion chamber. This can be important; every CC carved out means less compression ratio potential is available.
Now that we are clear that everything in terms of flow starts at the valve seat, let’s move on to the subject of the ports themselves. Although a flow bench is a prime requirement for optimal results, there are many ground rules that can be applied to improve a typical stock cylinder head, be it a two- or four-valve design.
Let’s start with the basic requirements of a port. Assuming the port is supplying a good valve-seat design, its first priority is to flow as much air as possible. The second general rule here is that whatever airflow is achieved must be done so by highflow efficiencies, as the port must not have an overly large cross section. If the cross section is too large, the port velocity is low, cylinder ramming decreases, and flow reversion increases. The result is poor lowspeed output with possibly no benefits at the top end either.
Another factor that can greatly affect low- and, to a lesser extent, high-speed output is swirl- or otherwise-induced mixture motion within the closed cylinder. In addition, it helps to make sure no big problems exist with fuel management within the port. This comes under the heading of “wet flow.”
Finally, we have to look at the combustion chamber form as defined by both the combustion chamber and the piston crown. That, in simple terms, means getting the required CR without producing a poor-burning chamber in the process.
The most basic port we could have is a round one that has a bend in it to accommodate the valve. Using this as a starting point, we can develop a port using some simple logic. Illustration 6-10 shows our starting point.
With any port, the radius of the short side turn is usually the numberone obstacle to achieving good midto high-lift flow figures. Formula 1 engines have a very large short-side turn radius and the port’s downdraft angle is less than 30 degrees off vertical. This makes for a very simple port that requires very little in the way of Band-Aid fixes to make it work extremely well. Unfortunately, the dictates of less-than-ultimate power on a less-than-ultimate budget and the low hood lines of street vehicles mean ports that are, without doubt, severely compromised. Illustration 6-11 shows the basic steps that were taken to make the port more efficient.
If you are modifying a typical parallel-valve two-valve head, then the last step is an important one to keep in mind. Understanding that the port more than likely needs a bias is the key to getting those big high-lift flow numbers. This leads ultimately to high-lift flow efficiency figures that exceed those delivered by a four-valve head.
The recovering flow efficiency of a parallel two-valve head is one of the reasons why this type of head responds to high valve lift, so bear that in mind when it comes time to spec-out the cam and valvetrain. The lift to shoot for is 0.30D for a hot street machine and as much as 0.35D for an all-out racer.
The optimal cross-sectional area for a given size of cylinder can float around somewhat, depending on the size of the intake valve being used, how tortuous the basic port is, and how big the cylinder may be. A good starting point for the intake is to have a nearly parallel section of the port, about 1.5 to 2 intake valve diameters up from the intake valve, sized to a cross section equal to 77 to 80 percent of the area of the intake valve itself. This area needs to extend, if necessary, into the intake manifold for about 2 to 3 times the diameter of the intake valve for a two-valve head and about 4 to 5 if we are dealing with a four-valve head. For the record, it is better for the widest powerband and best torque if one errs on the smaller side; this produces a punchier driving experience. Making the port too big can hurt output everywhere. Chart 6-12 shows the test results run on a small-block Chevy.
Applied Basic Porting
Now that we have plowed through the basics in theory, you might be asking, “What’s it worth in terms of power?” This obviously varies from head to head and engine to engine. The better the head is to start with, the less difference your efforts from a basic porting job will be.
Let’s use a middle-of-the-road example, a 5.0 Ford Mustang V-8. The results shown in Charts 6-13 and 6-14 can be achieved if we take those heads and do nothing other than skinny-down the guide bosses, so they are only about 1/8 to 3/16 inch wider than the guide bore, and round off the short side turn. All this can beachieved in just a weekend of work. If you already have a die grinder and suitable carbides for all the other jobs that these tools come in handy for, your cost is virtually zero.
Regular or dry-flow testing is something most performance enthusiasts have at least heard of, but maybe not wet-flow testing. The purpose of wet-flow testing is to establish that the fuel and air arriving at the cylinder is still suitably well mixed. In technical terms, this is called mixture quality. If the fuel separates from the air more than a minimal amount, the power takes a dive and fuel consumption climbs.
Wet-flow testing is not so much about measuring what is happening. It’s about observing and using judgment and experience to ascertain just how good (or not) the mixture quality arriving at and entering into the cylinder is. Wet-flow testing is still a science/art practiced by few, but the few who do it often see some relatively dramatic returns on time invested. My own experience investigating mixture quality arriving at the cylinders in the 1970s proved its worth by netting a couple of major championship wins. It can be a complex subject, and worth far more than the few pages I devote to it here.
The preparation for a quality mixture starts at the fuel injector or the carb booster. At part throttle, the manifold vacuum can almost completely vaporize the fuel. That’s good for combustibility and is a key ingredient to making good mileage. However, at WOT, there is not much vacuum to assist with mixture preparation. The fact is that, for the most part, the situation deteriorates a short distance downstream of the injector or carb boosters.
Our initial move here is to make sure we are using an injector or booster that does a good job. Injectors that are working properly are usually close to as good as they can be. The same cannot be said of carb boosters; so if mixture preparation appears to be a problem, investigate higher-atomization boosters. Having to run a richer-than-expected mixture to get maximum power is an indicator here, along with excessive fuel wash, exhaust smoke, and poor WOT fuel consumption. Steps must be taken to remedy the situation. The wet-flow bench points the way, but most fixes are less than obvious. But there are some moves that can be readily appreciated. The intake port “shear dam” is a prime example of this and has been very successfully used in several highly developed heads.
A lot of what it takes to get good mixture quality at the time the plug fires is about velocity. Good velocity through the port is the place where it all starts. Having good port velocity allows us to suitably direct that velocity to generate swirl (in the case of a two-valve engine) or tumble (in the case of a four-valve). Also, mixture motion produced by the quench of squish action is key. Burning the charge faster makes for not only more effective combustion but also reduces the chance of detonation.
If you need to lower the compression of an engine to have it compatible with a supercharger, do not just park the piston farther down the bore. In most instances, this serves to make the engine more prone to detonation because the charge burns slower. In so doing, it has more time to radiate heat to the as-yetunburned part of the charge. At some point the combination of heat and pressure causes it to detonate. Having a functional quench action not only allows the engine, in most cases, to make more power, but also to do it on less octane.
So far we have talked of many ailments that an intake port can have in terms of its ability to negatively impact mixture quality. Along with this, the only fixes shown have been complex and require a lot of high-dollar test equipment. This begs the question as to what the guy who has a lot less in the way of flow and dyno equipment can do to make the best of the situation. Essentially, the simplest advice here is to start with a carb or injector system that delivers a suitably fine spray and feed this into ports that are neither too big nor too shiny. A good 60-grit emery finish in the ports and intake manifold runners can do much to stave off poor mixture quality.
Compression ratio is one of those commonly known engine factors that has big power implications, but is rarely understood. It has much influence on the so-called optimal engine component combination. So that we can better understand the issues involved, the first move is to establish exactly what the compression ratio is. Illustration 6-15 does just that.
This simple example shows how it works. Let us assume the cylinder has a volume of 100 cc and the combustion chamber volume (shown in red on the right hand cylinder) has 10 cc of volume. The CR is then the volume above the piston in the lefthand cylinder divided by the volume above the piston in the right-hand cylinder. The left-hand cylinder has 100 + 10 cc, while the right hand has just 10 cc. That equals 110 divided by 10—giving an answer of 11:1.
As the CR goes up, there is also an increase in torque output. Theoretically, this increase should occur throughout the RPM range, so power goes up correspondingly. Illustration 6-16 shows how output increases as the ratio from one point to another is increased.
We glibly talk of the compression ratio and understand that it is a function of the piston traveling up the bore on the engine’s compression stroke. We need to appreciate just how significant the compression ratio is to our ability to spec-out an optimal engine parts combination. And therefore we must consider what happens as the piston comes down from top dead center (TDC), on the power stroke, in terms of the expansion ratio.
The fact that increasing the CR is good for output does not give us a license to just dive in and crank that number up without giving due consideration to the consequences. As the CR is increased, so the heat of compression and combustion temperatures increases. This makes it harder for the air/fuel mix to resist spontaneous ignition (detonation). When detonation occurs, the resultant super-rapid pressure rises and temperatures rise to where they quickly destroy pistons. Just how much compression can be used depends on many factors, with the most commonly appreciated one being the fuel’s octane value—the higher the octane, the greater the fuel’s resistance to detonation.
Regardless of the fuel’s octane value, we should strive to achieve a combustion cavity that is as resistant to detonation as possible. This means producing a compact shape that has good mixture motion within and so burns rapidly. It is easy to get carried away with thoughts of a big compression number and overlook the chamber that may result. Keep in mind that a charge trapped in the quench area does not burn well, if at all, until the gap between piston and head exceeds about 0.100 inch. Also, big piston domes do not help achieve an effective burn. From personal experience, I can say that a piston crown shape that inhibits rapid and complete combustion can wipe out 10 percent of the engine’s potential output.
When building an engine, it is best to machine the head to minimize the combustion-chamber volume first, and then use a piston with the appropriate crown shape to get the desired CR. In almost all cases, shallow-dish or flat-top pistons are best. However, a well-thought-out raised crown for those higher ratios (12:1 on up) can pay off big time, but it may take dyno testing to come up with what is best for the job.
CR—How High Can We Go?
If we put an engine on the dyno and keep cranking up the CR, the gains produced taper off until, at about 17:1, no further gains are seen. The increase from about 14:1 to 16:1 is very minimal, so why would we want to go up to these super-high ratios in the first place? This is a good question, and it is a classic case of getting the right spec/parts combination that ultimately produces the winning results.
The CR affects several important and power-influencing factors. Understanding these will put you more than one step ahead of the competition, so let’s investigate. The two most influential factors the CR affects are: the balance of the intake-to-exhaust-valve size, and just how much cam the engine can successfully use.
It is important to understand that the valves for a high-performance engine need to be as big as possible, bearing in mind any mechanical constraints. For a twovalve engine, this results in valve proportions in relation to the bore as per Illustration 6-16. Having established that cramming in the most valve area possible is a requirement, we have to ask ourselves how much we should apportion to each valve. The usually touted figure here is that the exhaust valve (assuming the same intake-to-exhaust flow efficiency) should be 75 percent of the intake valve. Although some engine builders stick to this number as if it were immutable, the fact is it floats in relation to the CR used. To see how this works, we need to move from considering the compression ratio to considering the expansion ratio.
Take a good look at Illustration 6-17. First, let us deal with how the CR (and by inference, the expansion ratio) affects the valve-size proportions. Let us assume two cylinders— one with a 15:1 CR and the other with a 2:1 CR. While they are at TDC, let us fill both cylinders with compressed air to 1,000 psi. If we now let the pistons go down the bore, the pressure for the 2:1 cylinder decays as per the darker of the blue lines, and the pressure for the 15:1 decays as per the red line. Note how the pressure drops far faster for the higher compression.
Next, let us say that we are going to open the exhaust valve 30 degrees before top dead center (BTDC). Doing so with the 15:1 CR cylinder means that the pressure has dropped to about 10 percent of the original; so 90 percent of the pressure energy has been used to turn the crank. If the exhaust valve is opened at the same point of the 2:1 cylinder, the amount of pressure energy remaining and subsequently wasted is about 30 percent.
The inference here is that a highcompression cylinder does most of its useful work earlier in the expansion cycle. Given a high enough CR, opening the exhaust valve earlier than normal has little affect on the power produced by the cylinder. On the other hand, opening the valve early on a low-compression cylinder throws far more pressure energy away. For a high-compression cylinder, the consequences are that the exhaust valve can be opened sooner and therefore does not need to be as big. Making it smaller means more room for the intake valve; so the ratio of exhaust to intake for, say, a 17:1 Pro Stocker, is going to be about 65 percent.
Conversely, a low-compression engine needs to hold the exhaust valve on the seat longer to make the most of the cylinder pressure. This means it has to be bigger so that, when opened, it has blown down sufficiently to minimize exhaust stroke pumping losses. Here, we can be talking of an exhaust-to-intake ratio of 80 to as much as 85 percent. Because of its inherently lower CR, a supercharged engine needs to have a bigger exhaust in relation to the intake than does a non-supercharged engine.
The fact that a high-compression cylinder does its work earlier in the cycle also means that a shorter rod can be used without such a big sideloading friction penalty. You can see how the CR affects the optimal rod length and the ratio of intake to exhaust. There are few one-size-fits-all parts or specifications within an optimal high-performance engine. This explanation of how the CR affects valve sizes and rod lengths for best output also demonstrates why changing just one thing on the dyno to see if it works is not necessarily the thing to do. It also demonstrates why assuming one aspect needs to be a certain value, come what may, can be a mistake.
In practice, a 2:1 cylinder does not achieve peak pressures as high as those seen by a 15:1 cylinder. Peak pressures increase as the CR goes up and a good general rule here is that, for a street engine, the peak is 100 times that of the ratio involved (i.e., 10:1 produces 1,000 psi). For a race engine, where volumetric efficiencies can exceed 100 percent, the peak pressures can be as much as 120 times that of the CR. Again, take a look at Illustration 6-17 and compare the red line with the light-blue line. These, on the same scale, represent the starting and finishing pressures of a 2:1 cylinder compared to a 15:1 cylinder. The shaded green area between the two shows the difference in the available pressure to turn the crank.
Cam Specs and Dynamic CR
One way of boosting power is to throw in a long-duration cam, but the down side is that such a cam favors top-end output at the expense of low-end output. Worse yet, it affects the CR while the engine is running. This is known as the dynamic CR, as opposed to the static CR discussed above. If the cam opened and closed the valves at TDC and BDC, then both these CRs would be the same. However, valve events are spread such that the valves open sooner and close later than the TDC and BDC positions.
Since intake valve closure is delayed, the piston has moved up the bore somewhat before the valve closes and traps the charge. This means that, at low speed, the CR that is actually seen by the cylinder is somewhat lower than the calculated static CR. If this results in too low of a CR, lowspeed output suffers accordingly. Initially, a certain amount of intakevalve-closure delay (about 30 degrees after BDC) has little effect on the dynamic compression because the crank/rod geometry is such that initially the piston moves very little up the bore, but that state of affairs does not continue for long. A good working rule here is that (assuming the cylinders seal up as they should) the cranking pressure for a street engine running premium fuel is about 190 to 200 psi. If the engine is below that, then there are ft-lbs still to be had, and above—you might want to use an octane booster of some sort.
Big cams need higher compression to not only allow the production of the best torque and top-end output but also low-speed output. Choosing a cam that is too long for the CR is usually a good way to lose low speed and gain little or nothing at the top end. Being a little more conservative on a cam selection can actually result in more power everywhere.
There are other factors involved with the compression ratio used and the cam spec for optimal output, and these are discussed in Chapter 10.
Written by David Vizard and Posted with Permission of CarTechBooks