Porting Principals - EFI Dyno Tuning LC



Porting Principals

Parts of the Port

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Parts of the port and their terminology

Areas of Importance

Considering the flow through the intake port as a whole, the greatest loss must be downstream of the valve due to the lack of pressure recovery (or diffusion). This loss is unavoidable on intake ports due to the nature of the poppet valve. On the exhaust ports the opposite condition exists and we are able to control the geometry down stream of the highest speed section, namely the valve seat. This allows the possibility of good pressure recovery and is the reason exhaust ports flow better than intake ports of equal size do.

Accepting the expansion into the cylinder loss as unavoidable, the rest of the port becomes that much more important. The areas which pass the most air at the highest speed for the longest time are the areas that are most important.

The valve seat configuration on the port and on the valve together form one of the most critical areas in the port. The highest speed seen in the port will be at or near the valve seat for most if not the entire duration of the cycle. After that the throat area and short turn radius become critical at higher lifts in the middle of the cycle. The valve seat and valve head angles should be studied carefully in each design.

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Sometimes in the pursuit of airflow, greed can get the best of any porter, and the tendency is to go too big in some places. Nowhere is the price to pay higher than going too big in the port throat, the point of constriction just below the valve seat. Make the throat too big, and the venturi effect is ruined, and usually the flow will be too. Keep the intake port throat no larger than 90percent of the valve diameter, and the exhaust throat down around 85percent.

You do not want the throat too big in relation to the rest of the bowl. Bowl hogs usually do this. You want the same or slightly larger cross sectional area at the pushrod restriction as the throat area. Over the short side will be even larger. Low lift cams (.550 and below) will not want the runner ground with equal cross sections at the runner throat whereas cams with high lift will. Smaller lift cams will want to be smaller in section to keep velocity up since the lift is short and the valve is not moving as much air. Basically, with high valve lift, the pushrod area can become a choke point whereas with low lift it usually won’t, unless it is extremely small.

The bowl area and the rest of the length of the port have important functions in controlling some of the dynamic behavior of the waves that traverse the system as well as setting up the air for a good entry to the throat. Shape, cross section, volume, cylinder swirl or tumble and surface finish are factors which must be considered in concert with the overall design of the rest of the engine and vehicle to achieve good results.

Zeroing Out Geometric Shrouding.

When addressing valve shrouding with the intent of minimizing it we need to make a start somewhere and ascertaining what the form of a chamber may be, if it was geometrically un-shrouded, is as good a place to start as any.

The breathing area presented to the chamber by a valve moving through its lift envelop is not quite as simple a geometry problem as it may first appear. The reality is that as the valve lifts it moves through three distinct regimes, each of which requires its own particular set of math formulas to produce an answer as to what the through-flow area is. We are not going to deal with this now as it is more advanced stuff. However, even if we ignore that we can still come up with a very good approximation of what it takes in the way of chamber form to produce a geometrically un-shrouded chamber. What we find is that at low lift the angle of the chamber wall as it leaves the valve seat needs to be very close to 45 degrees and as the lift progresses up to the critical 0.25 D lift point the angle needs to increase to about 52 degrees from horizontal.

The drawing below gives us a good guide to the form that needs to exist around a valve as it progresses through its lift envelope to ensure that the flow area around it is at least equal to the effective curtain area beneath the valve head.

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Look closely at this drawing. The green line represents the angle of the chamber wall as it comes off the seat. For all practical purposes this is right around 45 degrees. As the valve lift progress the point of zero shrouding of the edge of the valve in relation to the chamber wall gets slightly steeper until at 0.25D the wall angle is close to 38 degrees off the vertical (52 from horizontal) as represented by the blue line. Although not totally accurate we can say, within close limits, that when the valve is at 0.25D lift the gap between it and any possible obstruction should be equal to a minimum of 0.20D. Above 0.25 D valve lift the chamber wall can be vertical for zero geometric shrouding as the valve has reached the limit of the area it will present to the cylinder.

Wave Dynamics

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When the valve opens, the air doesn’t flow in, it decompresses into the low-pressure region. All the air on the upstream side of the moving disturbance boundary is completely isolated and unaffected by what happens on the downstream side of it. The air at the runner entrance does not move until the wave reaches all the way to the end. It is only then that the entire runner can begin to flow. Up until that point all that can happen is the higher pressure gas filling the volume of the runner decompresses or expands into the low-pressure region advancing up the runner. (Once the low pressure wave reaches the open end of the runner it reverses sign, the inrushing air forces a high pressure wave down the runner.)

Conversely the closing of the valve does not immediately stop flow at the runner entrance, which continues completely unaffected until the signal that the valve has closed reaches it. The closing valve causes a buildup of pressure which will travel up the runner as a positive wave. The runner entrance continues to flow at full speed, forcing the pressure to rise until the signal reaches the entrance. This very considerable pressure rise can be seen on the graph below. At the closing of the intake valve, pressure rises far above atmospheric.

It is this phenomenon that enables the so-called “ram tuning” to occur and it is what is being “tuned” by tuned intake and exhaust systems. The principal is the same as in the water hammer effect so well known to plumbers. The speed that the signal can travel is the speed of sound in the gas inside the runner. The boundary between the wave affected gas and unaffected gas could be compared to the event horizon of a black hole.

This is why port/runner volumes are so important. The volumes of successive parts of the port/runner control the flow during all transient periods. That is any time a change occurs in the cylinder whether positive or negative. Such as when the piston reaches maxumum speed half way down the stroke.

The wave/flow activity in a real engine is vastly more complex than this but the principle is the same.

At first glance this wave travel might seem to be blindingly fast and not very significant but a few calculations shows the opposite is true. In an intake runner at room temperature the sonic speed is about 1100 feet per second and will traverse a 12 inch port/runner in 0.9 milliseconds. The engine using this system, running at 8500 RPM, takes a very considerable 46 crank degrees before any signal from the cylinder can reach the runner end. 46 degrees during which nothing but the volume of the port/runner supplies the demands of the cylinder. This not only applies to the initial signal but to any and every change in the pressure or vacuum developed in the cylinder.

Why couldn’t we just use a shorter runner so the delay is not so great? The answer lies at the end of the cycle when that big long runner now continues to flow at full speed disregarding the rising pressure in the cylinder and providing pressure to the cylinder when it is needed most. The runner length also controls the timing of the returning waves and cannot be altered. A shorter runner would flow earlier but also would die earlier while returning the positive waves much too quickly and those waves would be weaker. The key is to find the optimum balance of all the factors for the engine requirements.

Further complicating the system is the fact that the piston dome, which is the source of the signal, continually moves. First moving down the cylinder, thus increasing the distance the signal must travel. Then moving back up at the end of the intake cycle when the valve is still open past BDC. The signals coming from the piston dome, after the initial runner flow has been established, must fight upstream against whatever velocity has been developed at that instant, further delaying the signal. The signals developed by the piston do not have a clean path up the runner either. Large portions of it will bounce off the rest of the combustion chamber and resonate inside the cylinder until an average pressure is reached. Then there are temperature variations due to the changing pressures and absorption from hot engine parts. These variations cause changes in the local sonic velocity.

When the valve closes, it causes a pile up of gas giving rise to a strong positive wave which must travel up the runner. The wave activity in the port/runner does not stop but continues to reverberate for some time. When the valve next opens, the remaining waves influence the next cycle.

[pic]The graph shows the intake runner pressure over 720 crank degrees of an engine with a 7-inch intake port/runner running at 4500 RPM, which is it's torque peak (close to maximum cylinder filling and BMEP for this engine). The two pressure traces are taken from the valve end (blue) and the runner entrance (red). The blue line rises sharply as the intake valve closes and this causes a pile up of air which becomes a positive wave reflected back up the runner and the red line shows that wave arriving at the runner entrance later. Note how the suction wave during cylinder filling is delayed even more by having to fight upstream against the inrushing air and the fact that the piston is further down the bore, increasing the distance.

The goal of tuning is to arrange the runners and valve timing so that there is a high-pressure wave in the port during the opening of the intake valve to get flow going quickly and then to have a second high pressure wave arrive just before valve closing in order to fill the cylinder as much as possible. The first wave will be what is left in the runner from the previous cycle while the second will primarily be one created during the current cycle by the suction wave changing sign at the runner entrance and arriving back at the valve in time for valve closing. The factors involved are often contradictory and requires a careful balancing act to work. When it does work, it is possible to see volumetric efficiencies of 140%, similar to that of a decent supercharger.

The "Porting and Polishing" Myth

It is popularly held that enlarging the ports to the maximum possible size and applying a mirror finish is what porting is. However that is not so. Some ports may be enlarged to their maximum possible size (in keeping with the highest level of aerodynamic efficiency) but those engines are highly developed very high speed units where the actual size of the ports has become a restriction. Often the size of the port is reduced to increase power. A mirror finish of the port does not provide the increase that intuition would suggest. In fact, within intake systems, the surface is usually deliberately textured to a degree of uniform roughness to encourage fuel deposited on the port walls to evaporate quickly. A rough surface on selected areas of the port may also alter flow by energizing the boundary layer, which can alter the flow path noticeably, possibly increasing flow. This is similar to what the dimples on a golf ball do. Flow bench testing shows that the difference between a mirror finished port and a rough textured port is typically less than 1%. The difference between a smooth to the touch port and an optically mirrored surface is not measurable by ordinary means. Exhaust ports may be smooth finished because of the dry gas flow but an optical finish is wasted effort and money.

The reason that polished ports are not advantageous from a flow standpoint is that at the interface between the metal wall and the air, the air speed is ZERO. This is due to the wetting action of the air and indeed all fluids. The first layer of molecules adheres to the wall and does not move significantly. The rest of the flow field must shear past which develops a velocity profile (or gradient) across the duct. In order for surface roughness to impact flow appreciably, the high spots must be high enough to protrude into the faster moving air toward the center. Only a very rough surface does this.

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The developed velocity profile, above in a duct, shows why polished surfaces have little effect on flow. The air speed at the wall interface is zero regardless of how smooth it is. Surface roughness (Reynolds Number) does have an affect on the velocity profile. Smoother walls produce long spikes in velocity and rough finished tend to keep the profile more compact.

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If you have one, get out your high school physics book and study up on Bernoulli's equation. It describes the relationship between pressure and velocity in a fluid as it flows through a pipe, which changes in cross sectional area along its length. Bernoulli's equation translated, says that as you increase the velocity of the fluid, the pressure of the fluid at that point decreases, and if you slow the fluid down, the pressure of the fluid increases, and how much it increases or decreases. How do you change the speed of the air in a port? Simply by making the port bigger (slower) or smaller (faster). Also, no fluid, including air, likes to change direction, because doing so causes it to lose velocity and energy which is hard to recover. To better understand your mission in porting heads or intakes, you should spend some time thinking about how all this works in an engine.

The cylinder head is the part of an engine that is most responsible for its performance characteristics. Once the basic geometry of an engine is established, there is no other part that has as much influence on the amount of power developed, and the shape of the power curve. All the other parts are merely supporting cast.

So, what determines the worth of one head over another? First, you must understand that any design is a compromise between what is desirable and what is possible. Engineers who initially design an engine rarely have free rein to make it the most powerful piece possible - and they may not want to either. Even in Formula 1 racing, where engines are designed from scratch to make as much power as possible, there are compromises that are determined by the rules of the sanctioning body and the necessity to install the engine in the car. Like other vehicles, aerodynamics and handling requirements require compromises in size, shape and weight of the engine.

The vast majority of today's popular aftermarket cylinder heads are compromised because they adhere to standard OEM port geometry. This is done so the supporting components designed to that geometry can be used on the new head. As engine builders, most of us have to work with parts that already exist. They may be production parts or aftermarket parts, but they all have compromises, and it's up to us as porters, to minimize the compromise.

How Airflow is Measured

Most people interested in performance know that a flow bench is used to measure airflow, but lacking hands-on experience, don't understand how it works, how it is used to measure flow in a cylinder head, or what the flow numbers actually mean.

You should know that a fluid flows from high pressure to low pressure. Air, being a fluid, follows that rule. If you turn on your vacuum cleaner, the motor creates a low pressure inside the cleaner, and atmospheric pressure, now being higher, pushes air in to fill the void. The rate of flow of the fluid is proportional to the difference in pressure. Seal the vacuum hose to the combustion chamber of a cylinder head, open the intake valve and turn the motor on. Air will flow through the port and into the cleaner. Add a valve in the hose to regulate how much pressure or suction you are using, and a means of measuring the suction and the amount of flow - usually done with manometers - and you have a flow bench.

All this flow bench does is move air through a port by creating a predetermined pressure differential, and then measure the quantity of air being moved. Tests can be done at any pressure you choose, up to the limit of the bench's capability. Most are done at 10", 25", or 28" of water, but the trend is to higher pressures, like 60".

A cylinder head adapter is commonly used to mount the head to the bench (as opposed to the vacuum hose previously mentioned) so the effect of cylinder wall shrouding can be simulated. Either a radiused inlet guide or an intake manifold can be attached to the head to eliminate turbulence at the manifold flange, and if testing the exhaust side, a short length of appropriately sized exhaust tubing is mounted on the header flange. A rigid fixture that will open the valve in .001" increments is needed too. Mount the head on the head adapter, open the intake valve to the first increment of lift, say, .100". Turn on the motor, and set the control valve at a test pressure such as 25", and record the amount of flow. Open the valve to the next increment of lift, such as .200", and repeat the test, again at 25" of water. A similar test is done at each increment of lift you wish to test. You have now flow tested one intake port, and have some data telling you how much flow your port has at 25" of water, at each increment of lift you tested. To be meaningful, all tests should be done at the same pressure, and use the same inlet or outlet configuration, and the same test procedure. Simply using a different inlet radius, valve shape, or cylinder diameter can change the flow. In other words, sweat the small stuff and pay attention to the details to ensure repeatability and make comparisons from one test to another valid.

To this, you can add tests for tumble and swirl in the cylinder, do localized testing within the port with test probes to determine velocity distribution and turbulent areas, try different valve and seat shapes, and even do wet flow testing if you have the capability. You can also reverse the head on the flow bench and check the flow characteristics around the valves in the combustion chamber (blow through the intake port and suck through the exhaust port).

A standard for maximum flow through a valve is 146 CFM per square inch of valve opening. This is used to rate the efficiency of a port. I use the valve curtain area - the circumference of the valve head (3.1416 x diameter), x valve lift, x 146 CFM, and then divide the result into the flow. This gives a percentage of the standard at each valve lift. If the port in the head can be made to flow up to the standard then it would in effect be achieving 100 percent efficiency. If it only flows half as much, then it would be 50 percent efficient. Another way to rate flow is to relate it to the area of the head of the valve. So now you have a means to test a port and a means to rate that port at each valve lift, relative to a standard. From that you can evaluate the efficiency of different heads based on valve size and lift.

Remember, a flow bench, like a dynamometer, is just a tool. It will give you data, but it will not tell you how to interpret that data, nor tell you what decisions you need to make regarding the suitability of the port you tested, nor will it tell you how to change the port to make it better. From this point, you must evaluate your data and make those decisions - and therein lies a large portion of the skill in modifying cylinder heads.

How Airflow Influences Engine Performance

Volumetric efficiency (VE) is the measure of how well the cylinder is being filled with air, as a percentage of what it would be if it were filled to the same pressure as the atmosphere outside the engine.

As the piston moves away from TDC, it creates a vacuum in the cylinder, which draws in the fresh charge of air and fuel. As the piston accelerates from TDC toward its point of maximum velocity at around 75 degrees after TDC, flow lags behind demand, creating a higher and higher vacuum in the cylinder. Then the piston slows down until at BDC it parks for a brief period before starting back up the cylinder. As the piston nears BDC, the inertia of the air from its velocity causes flow to catch up with piston demand and then finally exceed it.

After bottom dead center, the piston is going the wrong direction to pull in air, so even though we don't get 100 percent VE on the down stroke of the piston, we can achieve high overall volumetric efficiency by holding the intake valve open long after bottom dead center, to take advantage of the momentum of the intake charge (from its velocity) and the resonant tuning of the intake port, to keep filling the cylinder after the piston changes direction. This packing of the cylinder continues as the piston continues back up the cylinder until the valve closes. In a properly "tuned" intake system, a pressure wave will also arrive at the valve shortly before it closes, packing even more air in the cylinder. If it were not for the inertia of the incoming air, and resonant tuning of the port, it would be impossible to achieve even 100 percent cylinder filling. Typical low performance production engines operate in the 60 percent VE range.

The demand on the intake port is partially a function of the piston speed, which is zero at top and bottom dead centers, and can reach 8,000 feet per second or more (about 5,400 mph) somewhere between 70 and 80 degrees after top dead center. For a 4" bore by 3.48" stroke 350 Chevy to achieve 100 percent volumetric efficiency at this point, the head would have to flow over 500 CFM. It is unlikely that we will find a head for a small block Chevy that will flow this much, and if we did size the port for the 500 CFM demand at this point in the cycle, it would be too big during the remainder of the cycle and have insufficient velocity to continue filling the cylinder after bottom dead center. Ports and valves that are too big don't generate sufficient velocity in this segment of the intake cycle, and therefore don't make as much power as a port that is properly sized. Bigger is not always better. With resonant tuning and high velocities, typical performance engines can operate close to 100 percent, while more highly refined engines can achieve 120 percent or more, especially engines with multiple intake valves per cylinder.

On the exhaust side, velocity is important too. The velocity of the exhaust pulse in the port and header creates a vacuum behind it, creating a pressure drop in the cylinder as the piston approaches TDC on the exhaust stroke. This pressure drop from the exhaust during valve overlap, gets the intake system started before the piston even starts down on the intake stroke.

The kinetic energy in the air moving in and out of the engine is a function of the square of the velocity, so small changes in velocity make large changes in the energy of the flow. Even though the conditions in a running engine are constantly changing throughout each intake and exhaust cycle, steady flow on a flow bench can give a good representation of the power available from the engine by approximating the average conditions in the engine. Tests done at 25" of water-test pressure seem to closely approximate the average conditions that exist in an engine.

As I mentioned earlier, there is a trend to testing at much higher pressures. Peak velocities in a port can be over 600 feet per second, but testing an intake port sized for high rpm power at 25" of water may only have 200 feet per second on the flow bench. It may be very efficient at that velocity, but have high levels of turbulence at twice that velocity. The only way to determine the worth of a port at the higher velocities is to test it at those velocities, requiring higher test pressures.

When you increase the efficiency of a port in the normal direction, you also increase the efficiency of the port in the wrong direction. This makes the engine more sensitive to camshaft changes, and to intake and exhaust system tuning. An intake manifold or exhaust system that worked just fine before, may not work very well after the heads are done, not because it's a bad piece, but because it's the wrong piece.

So, what is the process to determine the correct cylinder head or head modifications needed? Whether you are dealing with a street rod or a serious race effort, the first thing to do is determine how the engine will be used, the rpm range it will see, the engine specs, and a power goal. I regularly deal with customers that have never answered those questions, they just want their heads ported because they believe it will make more power, but until they are answered, you have no way to determine the optimum port and valve sizes, and flow requirements. If you want a pair of heads for a rock crawler that mostly runs just above idle, but you ported them the same as you would for a 7,000 rpm bracket racer, you would probably not be happy with the results. Once these things are known, you can proceed.

There are mathematical relationships that can predict the airflow requirements to reach a specific power - or predict the power potential of a known airflow - and the rpm at which peak power will be developed. This includes the intake manifold and carb or throttle body. These equations can be used to see if you are in the "neighborhood" of what you want to do. If the rest of the engine combination is complimentary, they will be pretty accurate. If you have the airflow desired, but the engine does not make the power it should, then you have another problem that needs fixing.

Horsepower per cylinder = .43 x airflow @ 10" of water, .275 x airflow @ 25" of water, or .26 x airflow at 28" of water. To find required airflow for a given horsepower, divide the horsepower per cylinder by .43, .275 or .26 respectively. 300/8/.26=144cfm 350/8/.26=168cfm 400/8/.26=192cfm

RPM at peak horsepower will be 2,000 divided by the displacement of one cylinder x airflow @ 10? of water. Use 1,267 @ 25" of water, or 1,196 @ 28" of water. The more airflow available to the cylinder, the higher the rpm required to reach peak horsepower.

The maximum practical intake valve size in a two valve, wedge chamber head is about .52 x the diameter of the bore. In a hemi type chamber, .57 is about the maximum.

How to Improve Airflow

The vast majority of porting work can be described as merely cleaning up an existing port and doing a valve job for improved performance for a street rod, boat, or for sportsman level racing. No matter what the application, an understanding of what is needed and a methodical approach to the modifications will usually result in a satisfactory level of performance. A flow bench is essential, because it's really easy to make things worse. It can be a commercial unit or home built, as long as it is repeatable.

The level of porting you wish to do will determine the approach taken. You can start with a good valve seat shape and blend it into the bowl under the valve and to the chamber and call it good, or you can get into a complete ongoing cylinder head development project, or somewhere in between.

The area from around the valve guide across the seat and into the chamber responds more to changes (both good and bad) than any other part of the port, especially the short turn, and the shape of the valve and valve seat is the single most critical part. The largest flow loss is in the expansion of the air as it exits the valve/valve seat into the chamber. Different valve shapes which sometimes include rolling or radiusing the margin area of the valve, different widths and angles on the back side of the valve, and different angles approaching and leaving the seat are some of the "tunable" things that make an otherwise ordinary job come to life. Many ports can be improved substantially by merely blending the seat cuts into the bowl under the valve and into the surrounding combustion chamber, frequently referred to as "pocket porting."

My approach to head work is to first determine a power goal as explained above, which determines the required flow and port and valve sizes.

Next, I flow the head to get a baseline flow curve, and decide what modifications are most likely needed to meet the performance goal. Depending on how much material has to come out, I may sonic test the ports for thickness to make sure I don't have a problem there.

If I'm to do a full porting job, I make a rubber mold of the port and slice the mold into segments 1/2" to 3/4" wide. I lay each segment on graph paper, draw around the circumference, and count the squares in the outline of each segment to obtain its cross sectional area. This gives me a diagram of the shape of the port. The silicone rubber I use is Dow Corning Silastic V base and curing agent. I have tried the Silastic M, but it is too hard, which makes it difficult to get out of the port once it cures. Expect to pay close to $150 for a gallon of it.

To the general automotive community, CNC porting is the hot item these days. CNC merely means that a computer controls the tool paths of a milling machine to take material out of a head. The main advantage to CNC porting heads is time and repeatability. In most cases, the port being CNC'd is a digitization of a port that was developed by hand porting using the methods described above, and the success of the port being CNC'd depends on how good the port is that was used to develop the program in the first place. If the original port wasn't too good, then all you have is a whole lot more ports that are also not too good. Even if it's an outstanding port, it may be wrong for your application. The key here is to find a good port that fits the application, and the only sure way I know to do that is to get your hands on one, make a mold to evaluate the shape, and flow test it, then decide if it's suitable.

Ken Weber, who formerly operated a marine engine rebuilding business, is an independent technical writer near Denver, CO, covering the high performance engine market.

Cylinder Head Tech:

At first, the task of clearing and recharging the cylinders in a high-speed, four-stroke engine seems impossible. Such processes need time, and it's hard to believe there's enough available for this one, which faces many impediments and is crowded into the merest fragment of a clock's tick.

The intake stroke lasts for 180 degrees of crank rotation, which is only three-thousandths of a second at 10,000 rpm. Camera shutter openings are as brief., but light has no mass and moves at 950 million feet per second. Air's mass makes it lag, and it hits a sonic wall about 1100 feet/second, with localized shock waves further blocking the intake ports at much lower air speeds.

Yet cylinders get filled-with efficiencies sometimes exceeding 100 percent-without mechanical supercharging. This is possible because the intake process actually begins in the preceding exhaust stroke and extends far into the following compression stroke. We've methodically learned to make the pesky effects of inertia work for us; and minimized the bad effects of problems that cannot yet entirely be solved.

On a cylinder head's intake side you have only atmospheric pressure, 14.7 pounds per square inch at sea level, working to stuff air into the cylinder. No matter how hard the descending piston tries it can't pull air in behind it. It can only create a space for atmospheric pressure to fill.

It's a different story over on the outlet side, where a pressure close to six atmospheres exists when the exhaust valve opens to begin the event called "blow down". Further, after blow-down, pistons mechanically force exhaust products from the cylinders, and do so against the resistance of undersized valves, badly designed headers or steel cork mufflers.

The more important exhaust event is the high-velocity shove the rising piston gives exhaust gases during the exhaust stroke. The shove peaks at maximum piston speed (in most engines occurring a little less than 80 degrees of crank rotation before the piston reaches top dead center), where it suddenly gets yanked to a stop. But the momentum of the gases in the exhaust pipe continues, leaving behind a partial vacuum. This starts the air/fuel mix above the part-open intake valve moving into the cylinder before the piston begins its intake stroke.

Engines benefit from exhaust-augmented intake flow in two ways; an obvious advantage is that it gives the too-brief intake period an early start. The second effect, less obvious but also important, is that combustion chamber cross-flow during valve opening overlap (the period during which both intake and exhaust valves are open) clears residual exhaust gases, which slow combustion, depress power by displacing part of the fresh charge, and can require some weird kinks in the ignition advance curve.

Exhaust systems primarily aid intake flow by their manipulation of the combustion "sound wave". A sound wave creates a disturbance ahead of it and leaves one behind; such "positive" waves bursting from the exhaust port are followed by negative pressures. When the strongly-positive exhaust wave emerges from the end of a pipe, it leaves behind a negative-pressure tail, which then reflects back toward the port. If the length of the pipe is right, the negative wave will arrive back at the exhaust valve as the piston reaches TDC, thus further assisting in clearing the combustion chamber.

Sound waves are reflected by any cross-section change in the duct in which they are traveling. The sawed-off end of a pipe is one such change; the closed end of a pie is another. The difference is that increases in section invert the wave while reflecting it, changing positive waves to negative and vice-versa; section reductions reflect the wave with the same sign.

While speaking of sonic waves, I should caution you about confusing their behavior with that of the media in which they travel. Like all sound-conducting media, air has mass and the other properties of matter. sonic waves are by contrast, purely energy and thus follow an entirely different set of rules. such waves make zero-radius 180 degree turns and reversals without delay or loss of strength.

` Plain pipe ends do a poor job of returning the energy of an emerging sound wave, which is why horns have flared open end-to get better energy recovery and thus amplitude. Megaphones, the exhaust pipe horns known in engineering as diffusers, are vastly more efficient in this regard. Racing two-stroke engines expansion chamber exhaust systems have elaborate blow-down diffusers, because of their heavy reliance on this vacuum-cleaner effect to pull air through the transfer ports.

Four-stroke engines seem perfectly happy running with plain parallel-wall pips, though engines developed for megaphones have to be reworked to function well without them. Harley-Davidson's famous racing chief, Dick O'Brien, never was totally convinced that the megaphones used on the "low Boy" KR's did anything but make noise. At the time I was sure he was missing something, but now I believe his reservations were valid.

Oddly, the 45-degree cut-off at the end of KR straight pipes did coax a tad more power out of H-D's cranky old side-valve engine; O'Brien was at a loss to explain this oddity. I tried a 90 degree cutoff once, and found the KR didn't like it. No coherent theory I've heard or conceived explains why that should have been so.

It now appears exhaust pipe diameter, meaning gas velocity in the exhaust system, is more important than sonic wave activity. actual gas velocities vary in ways tough to grasp and impossible to calculate, but the nominal speed is easy to figure and provides a useful rule-of-thumb: simply multiply piston speed by the ratio of cylinder bore and pipe areas.

Nominal gas speed were well below 200 feet/second in most vintage bikes, but in the AJS 7R of the 50's it was up to 220 feet/second. By 1972 the small diameter pipes on H-D's XR750 raised that engine's exhaust velocity to just above 300 feet/sec. The Triumph 650 TT Special I used to set a Bonneville record (and acquire an abiding dislike of Wendover, Utah) years ago also had small pies and 300-plus exhaust gas speeds. It had 1 3/8-inch pipes, which almost everyone thought too small. My slide rule said they were the right size, and the larger-diameter pipes we tried slowed the bike.

Gas velocity is even more important over on the engines intake side, where it packs air into the cylinder between the intake stroke's ending and intake valve closing. This is crucial, since with high-speed engines there is a significant lag between the piston beginning the intake stroke and the flow of air into the cylinder. Outflow in the exhaust can pull air across from the intake to give the intake process a head start, but cylinder pressure still precipitously falls through the first half of the intake stroke. Air simply can't keep up with the piston, which at 9000 rpm in the XR750 goes from it's stop at TDC to 80 miles per hour in 1.5 inches, reaching that speed in 0.0014 seconds.

Fortunately, the air inertia that delays air/fuel inflow causes it to crown in at the end of the intake stroke, and beyond. The XR750's intake ports are small enough to raise the nominal gas speed to 370 feet/second, which gives it plenty of momentum. This is why intake valve closing is delayed for many degrees after the piston has finished it's intake stroke and begun compression. Closing the intake valve while air is still flowing into the cylinder, or closing it after flow reverses, gives less than the best power. You have to close the intake valve(s) just as the inflow slows to a stop, thus trapping the greatest weight of air/fuel mixture in the cylinder.

Serious tuners need some means of shifting cam timing ( in increments no coarser than 1.5 degrees) to let them experiment their way to the optimum intake closing. This is usually done with multiple oversize bolt hoes in the driven cam sprockets and offset bushings, although my old Aermacchi required woodruff keys with a sideways-jog at the shaft and timing gear join to shift camshaft phasing.

High-performance engines' intake valves close typically 60 to 80 degrees after the intake stroke ends and the compression stroke begins, so you know gas inertia is playing a major role in cylinder filling; if it didn't there'd be no need to delay intake closing, and no sensitivity to the timing of that event. None of the other valve actions-exhaust opening or closing, or intake opening-are nearly as important.

Flow benches can be used to blow a lot of smoke up your shop coat when you're looking for horsepower. You can always make air flow numbers rise by increasing valve head diameter, or by enlarging the passages leading from the atmosphere. But higher air flow numbers do not necessarily translate into more power, as many in the engine development field (including yours truly) have discovered.

Mercedes-Benz made the big-port mistake with the design of its awesomely complex eight-cylinder M196 GP car, which had desmo valve actuation and intake ports the size of drains. They found themselves being out-horsepowered by the British Vanwall, with an engine that was virtually four Norton 30M Manx Cylinders and heads bolted to an aluminum Rolls Royce armored car crankcase.

Ford's 1960's four-cam V-8 also had huge intake ports, and while it turned more revs than the Offy four-banger engines then dominant at Indianapolis, it was no better than a match for them. When given an early peek at the Indy Ford's cylinder-head castings, I expressed the thought that its ports might be too big. Ford's engineers were too polite to tell me how absurd they considered my remark to be, but their expressions made it plain. I was too polite to send them an "I told you so" note after Dan Gurney sent one of the engines to Weslake Engineering in England, where it's intake ports were made smaller and its output got bigger.

Ford's engineers were then vastly ignorant of the world beyond Michigan's borders. They had no idea Harry Weslake and Wally Hassan (who created the very successful Coventry-Climax racing engines) had learned years before not to take too literally what the flow bench said. They were narrowing intake ports to provide nominal gas speeds in the range of 350 to 400 feet-second, making good use of the fact that kinetic energy packing air into the cylinders increases with the square of it's velocity.

To estimate runner air speed, take flow cfm @ 28” divide it by the limiting cross section area and multiply by 2.4, i.e.:

272cfm/2.1sqin*2.4 = 311 ft/sec theoretical. Actual is often higher.

FPS = ( CFM / CA ) * 2.4

CFM = FPS * CA * .41666667

CA = ( CFM / FPS ) * 2.4

RPM = ( FPS * CA ) / ( Bore * Bore * Stroke * .00353 )

FPS = ( Bore * Bore * Stroke * RPM * .00353 ) / CA

CA = ( Bore * Bore * Stroke * RPM * .00353 ) / FPS

Mach number = FPS / 1116

(.627 Mach = 127.5 % Volumetric Efficiency potential,.55 MACH = 121.1 % VE)

where;

RPM = point of desired Peak HP

FPS = Feet per Second

CA = Cross-Sectional Area in Square Inches (smallest measured)

614 fps = ~ .55 Mach. This is a flow rate that should not be exceeded due to flow difficulties.

============================================

use

RPM = ( FPS * CA ) / ( Bore * Bore * Stroke * .00353 )

i.e. if your 347 engine chokes @ 6600 rpm with a 2.2 CA

rpm = ( 614 * 2.2 ) / ( 4.030 * 4.030 * 3.480 * .00353 ) = 6771 rpm

pretty close to the 6600 RPM "Choke ?" you are experiencing

then solve the other way

CA = ( Bore * Bore * Stroke * RPM * .00353 ) / FPS

2.145 = ( 4.030 * 4.030 * 3.480 * 6600 * .00353 ) / 614

the 2.145 rounds off to 2.2 ..pretty close to your 2.2

So, what are the flow rates that we should shoot for – in general the following can be used as guide:

240 ft/sec - intake - ram effect faint (.21 Mach)

- exhaust- scavenge faint

260 ft/sec - intake - ram effect moderate (.23 Mach)

- exhaust- scavenge weak to moderate

280 ft/sec - intake - substantial ram (.25 Mach)

- exhaust - scavenge moderate

300 ft/sec - intake - * ideal ram (.269 Mach)

- exhaust - substantial scavenge

320 ft/sec - intake - possible loss (.287 Mach)

- exhaust - * ideal scavenge

340 ft/sec - intake - likely loss (.305 Mach)

- exhaust - possible loss

In practice rules of thumb have developed saying that at peak power, the ft/sec figure should be somewhere within 280-380 exh and 240-355 intake. Velocity over 600fps (.55 Mach) often cause inertia blocks and/or flow separation and take a very well designed port to work.

you can use this with Air Velocity FPS to solve for what is the required Intake Valve diameter needed for a certain "Peak HP RPM"

Intake Valve = (( RPM * CID ) / ( Cylinders * 314. 5 * 282.743)) ^.5

1. 528 = ((5,500*302) /(8*314.5*282.743)) ^.5

1.724 = ((7,000*302) / (8*314. 5*282.743)) ^.5

1. 60 = ((5,500*331) / (8*314. 5*282.743)) ^.5

1. 80 = ((7,000*331) / (8*314. 5*282.743)) ^.5

where;

RPM = the point you want Peak HP to occur

CID = total engine size in Cubic Inches

Cylinders= the number of engine cylinders

314.5 = Air velocity in Feet per Second

282.743 = Units Constant

^ .5 = Square Root of a Number

Discharge Coefficient

Quote from Darin Morgin

The "Discharge Coefficient" is the measure of how efficient a given area is in regards to mass flow verses area, divided by a theoretical maximum. I use the 146 cfm/sqin and not the 137 cfm/sqin that the SAE dictated years ago just because that's what all my data has been accumulated with from day one.

Window Area = Valve diameter * Pi * lift (Also called Curtain Area)

window area * 146 = theoretical maximum flow for that area

Take your flow and divide it by your theoretical maximum. This is the ratio of effective flow area to actual flow area - this is your discharge coefficient.

Cosworth Engineering's Keith Duckworth was the creator of the modern high-output four-stroke. Casting aside tradition, Duckworth combined large-bore short-stroke cylinders with narrow-angle valves and a compact combustion chamber. He didn't originate the use of high-intake port velocities to ram-charge cylinders, but he and those he's influenced now design for nominal intake speeds approaching 450 feet/second.

Of course, there's a lot more to cylinder gas exchange than port velocity. But unless you've spent eons dragging air through ports, manifolds, etc.,, at a flow bench, you probably have no real understanding of what aids flow and what slows it. If there is any rule for the inexperienced to keep in mind. it is that everything a reasonable intelligent person should intuitively believe to be right will probably be totally wrong.

Take valve shape for example, these days typically an un-streamlined disc on the end of a stick Your eye will tell you the shape is horrible, an example of how we've fallen into decadence since the days of those British power plants with beautiful, deeply tuliped intake valve. Then you hit the flow bench and find that the one with all the loveliness of an overgrown nail better at all lifts. And then you repeat the experiment with another port and find it responds better to a tuliped valve. Some ports are like that, by virtue of slightly different interior contours or different valve angles.

Or you can try valve seating surfaces-maybe someday you can tell me why sharp edges are better here than rounded ones. The worst valve I ever tested was one I made the mistaken belief my eye could judge how air would behave between the valve and seat. I ground a valve head with a radius instead of a flat where it seated, along with a similar-shaped grinding stone for the seat. Testing this idea required tons of work, yet my streamlined valve and seat combination was worse at all lifts than the typical series of abrupt, sharp-edged flats.

You'd think that getting the valve completely out of the way while flow-testing ports would let the air really whistle on through. But peak flow almost always occurs with the valve in place, at a lift equal to about 30 percent of valve diameter. And this is with a manifold and carburetor in place, and a cylinder between head and flow bench receiver ( the cylinder's adjacent walls can significantly influence flow around intake valve heads).

Multiple valves ( more than two per cylinder) actually offer little or no real valve-area advantage. You can prove this to yourself by drawing circles representing valves inside a larger circle signifying the cylinder bore, Unless you fudge the whole thing with unrealistic provisions for valve seats, clearance around the valves, etc., the total for valve head areas is about the same for two, three or even five valve layouts. The benefit lies in the fact that total head area counts only at or near full lift: at lesser lifts, flow is largely limited by the valve seat ring area, really more a function of the total of valve circumferences than area. Viewed this way, multiple valve layouts are better, though only Yamaha has found any gain with more than four valves.

Air flow in ports takes paths totally unlike those you would normally envision, unless you happen to have an abundant knowledge of compressible fluid dynamics. In your imagination, air may move in orderly lines of travel, with particles marching along the roof of the port staying high, those on the floor staying low, and all traveling in neat, linear streams. The reality is a very different matter.

When flow in a duct (an intake port, for example) arrives at a bend, it loses any semblance of orderly behavior. Particles on the inside of the bend travel the shortest distance (offering the least resistance to flow), so they tend to maintain speed in the downward turn to the valve seat. But flow in the top of the port slows relative to the floor, creating a large velocity gradient. Pressure in a moving fluid varies inversely with it's speed, so the velocity gradient creates a lower pressure at the port floor than at it's roof. this differential causes air at the sides to move upward and the midstream air to move down, with the resulting flow stream made to divide into to contrarotating vortices where the port bends. Add to this the invisible "smoke ring" vortex forming beneath the opening intake valve and you have enough disorder to confound even the best of minds (or computers).

Port and valve configuration (both shapes and angles) can profoundly influence combustion efficiency as well. Jack Williams AJS 7R made it's best power with an intake port shape that compromised flow in favor of creating more combustion chamber swirl and redirecting incoming fuel droplets away from the cylinder walls. I am reliably informed that Keith Duckworth has settled on the intake valves leaned 15 degrees from the cylinder axis, and ports at 30 degrees from the valves in a similar trade-off between flow and combustion.

Intake flow influences combustion because both carburetors, and fuel-injection nozzles deliver fuel in liquid form. The best you can hope for is a fog of droplets small enough to stay suspended in the air while evaporating; big drops are centrifuged out of the air stream, splatting against the intake port and cylinder walls, which is bad for power, fuel efficiency and emissions. Fuel can't burn until it evaporates; if you have raw fuel still trying to burn when the exhaust valve opens, it goes out the pipe, wasting your money and polluting the air.

My experience (not the final word on anything even for me) is that the biggest improvement in flow from a change in port shape- with the least port enlargement and resulting velocity loss- is obtained by widening the port floor upstream from the valve seat. Air likes to take the most direct route, and the more you ease that route the better flow becomes. Shaving metal out of the lower sides of the ports bend (making a D-shaped cross-section, with the port floor on the flat side has in my tests shown big flow improvements in sharply bent ports.

Smoothing intake flow (thereby minimizing the turbulence of the main flow stream) is best accomplished by making sure the port's section area decreases all the way from the carb inlet to the bend above the valve seat. The small diameter, high-velocity section of the port needs only a slight convergence of 1.5 degrees included angle, which doesn't sound like much. But a 12 inch section of aluminum pipe taper-bored for a 1.5 inch inlet and a 1.498 inch outlet flows better than a parallel-wall pipe, and a lot better than air going from the cones' small end to it's beg end. Sound waves love a divergent duct, air flow does not.

I'm not convinced that polishing a port's interior surfaces to a mirror finish does anything but look good. The problem here is that while we know there's a degree of roughness beyond which flow suffers, we can't agree on the limit to which polishing helps. One those rare occasions when I do porting myself, I settle for a smooth but not polished finish. If I were in the head porting business like my long-time friend Jerry Branch, I'd put a spit shine inside the ports and combustion chamber, just as he does. The way Jerry does it, his customers never have to wonder if the ports are smooth enough.

Jerry has discovered that some ports flow better if he cuts tiny slots across the floor of the bend upstream from the valve. The slots apparently act as turbulence generators that energize the air and make it stick to the port floor, following the bend more closely. That's the theory anyway, though like so much we believe about port air flow, it's arguable because air hides is secrets behind a cloak if invisibility.

In time, we will know a lot more about the details of flow in and out of cylinder heads. For decades, researchers have used smoke, pinwheels, dye droplets, etc. in their attempts to see what air is doing. The water-anaolgy method, where water substitutes for air and flow is made visible with fine bubbles or aluminum particles, is still used in many labs. But the growth of mystery-dispelling technologies has recently brought doppler-laser metering and computer imaging to the field. Maybe one day soon we'll learn why the things a century of experience has taught us actually do work, and why others do not.

Expert Advice:

Joe Mondello, who’s name has long been synonymous with high-performance cylinder heads, said a lot of people who don’t really know what they’re doing jump into head porting and make big mistakes.

"They take out metal where they shouldn’t be taking out metal and end up with ports that are too big and don’t flow as well as they should. The shape of the port is far more critical than the overall size of the port," stated Mondello.

Mondello, who teaches the secrets of building, porting and flow testing high-performance cylinder heads at his Mondello Technical School in Paso Robles, CA, said he also sells special porting tools that are designed for every part of the cylinder head.

"When you’re doing the short-side radius of a port, you don’t want to take out too much metal. You just want it to be nice and smooth," instructed Mondello. "Trying to get around the short-side radius bend is difficult unless you use a cutter that’s designed for that purpose.

"When cleaning up the bowl area, blending alone won’t improve flow unless you also remove some metal to increase volume. Many people don’t do valve bowls properly. You have to blend everything from the base of the valve guide to the base of the primary valve seat, and then do a 3-angle valve job. Otherwise you’re just scratching the valve bowl and ports, and aren’t really gaining anything."

As for matching ports, Mondello said not to use gaskets as a guide because there’s too much variation in gaskets and most aftermarket gaskets have openings that are up to 1/8" larger than the port runners. If the port is enlarged to match the gasket, it can reduce air velocity and hurt performance.

"We teach port matching, not gasket matching. I pick the largest port, match all the others to it, then do all the work inside the port to maximize air flow around the pushrod tube turn because that’s where the biggest restriction is in the port," said Mondello.

"The largest gains in horsepower are found on the intake side by raising the roof of the port (the side closest to the valve cover) by .100" to .175". The amount of metal in the top of the intake manifold runner will determine how high you can raise the roof.

"On late-model Chevy Vortec heads, you don’t want to change the shape of the port much. The best advice here is to clean up and equalize the ports so they have the same height and width. On small-block heads, there’s a large pocket right below the rocker arm stud in the roof of the port. This should be filled in with epoxy to improve air flow. Doing that will give you an extra 15 cfm.

"On exhaust ports, if you tried to match the port to a header gasket you’d probably destroy the port. The secret of exhaust porting today is not how big the port is, but the shape of the port and the velocity of the exhaust flowing through it. We don’t even flow test exhaust ports anymore because most heads have plenty of flow capacity as is. All we care about is velocity and pressure.

"Nearly every single exhaust port today, except for Ford 302, 5.0L and 351 heads, are big enough. The only thing we do to enhance air flow is raise the roof of the port about 0.100", depending on the headers used. We don’t touch the floor of the exhaust port or the sides unless we have to get rid of a hook, seam or rough area in the casting," said Mondello. "Any time you start making the ports bigger on the exhaust side, you usually end up killing air flow in the head. I’m talking a reduction of 25 to 30 cfm. All you need to do is clean up the valve bowl, blend the short-side radius, and raise the roof slightly. Don’t touch the floor or walls."

Mondello explained that CNC machining and hand grinding are two different techniques for porting heads. "Everybody says CNC is the way to go. But you first need someone who can take a raw casting and rework it so it has good air velocity and flows well. Then you can digitize it and reproduce it with CNC tooling on other heads. There are a lot of CNC profiles being sold today, but I think most have some room for improvement. Additional hand grinding can usually pick up another 10 to 12 or more cfm."

As for polishing, Mondello said a smooth finish is great for exhaust ports, but a rougher finish flows better on the intake side. He recommends using 300- or 400-grit paper followed by a Cross Buff for polishing exhaust ports, and 50- or 60-grit paper for the intake ports. A slightly rough surface texture in the intake ports and intake manifold runners creates a boundary layer of air that keeps the rest of the air column flowing smoothly and quickly through the port.

Porting Lessons Learned the Hard Way

1. Ports that make power are designed with size and shape in mind, not the gross airflow at peak lift.

2. The blend area under the 45-degree cut on the exhaust valve seat is critical to good flow there.

3. There should just be a break at the mouth of the exhaust port, not the 35- or 38-degree cut most use. The sharp edge of the cut causes the air to sit there and shear in a turbulent state.

4. Big ports yield impressive flow numbers, but not necessarily quick-revving engines. Velocity is the key to making horsepower in a hurry, and smaller ports with big flow numbers have velocity.

5. The short side radius on the intake port is one of the most important areas in the cylinder head combination.

6. Most of the air/fuel mixture enters the combustion chamber across as little as 120 degrees of the intake valve.

7. The carbon coloration left in the combustion chamber after the engine has been run can tell you a lot about where the air/fuel mixture is ending up in the chamber. The optimum is an even coloration everywhere in the chamber.

8. Low-lift flow, around 0.200- and 0.300-inch lift, is critical to having an engine that revs strong through the power band. You can't see this on the dyno, but drivers always talk about a slow-revving engine as lazy or weak, and they don't like driving them.

9. A domed piston with a large-volume combustion chamber is not conducive to complete, even combustion. A dished piston, where the dish mirrors a small chamber, is ideal because as the piston and cylinder head squish together, the areas that aren't part of the chamber force the air/fuel mixture into one central combustion area. This makes power.

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