Porting Principals
Parts of the Port
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.
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.
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
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.
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.
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.
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 viceversa; 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 fourbanger 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 taperbored 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 slowrevving 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.