The flow bench is used to measure intake and exhaust port flow at various valve lifts. This data must be evaluated against the measured cross-sectional area of the ports. The larger the cross-sectional area of the intake port above the valve, the higher the engine must be revved to create sufficiently strong wave action in the intake runner. Best performance comes from high port flow with a minimum cross-sectional area. Worst performance comes from over-porting above the intake valve, even if the high-lift flow data looks acceptable. The flow bench results also reveal to the engine builder the flow bias between the intake and exhaust ports, and the low-lift flow to high-lift flow relationships. This is critical information for fine-tuning the camshaft design, and the header design. The exhaust port flow measurements should be taken with a port-matched header pipe bolted onto the cylinder head. The intake port flow measurements should be taken both with and without the intake manifold installed (no throttle assembly or other restriction to the runner entry). The throttle assembly with air cleaner should be flow benched separately.

For many applications, stock cylinder heads can be successfully used when combined with other carefully chosen engine components. Many factory cylinder heads show a substantial improvement in port flow when pocket ported. Pocket porting consists of minor blending and cleanup of sharp edges in the ports immediately above the valves. Leave the basic port shapes alone when pocket-porting stock heads, and get a high performance valve job. Then use the heads to their best advantage in the RPM range they were designed to operate. Properly designed headers and a properly designed camshaft are essential to making maximum power from an otherwise stock engine.

As a general recommendation for maximum horsepower racing applications, the exhaust port in your fully prepared cylinder head should flow about 75%-78% of the intake port flow rating at any given valve lift. This will allow the header to properly scavenge the cylinder toward the end of the exhaust stroke, which is essential for extending the RPM range of the engine well past the peak horsepower RPM. Headers designed using the Header Design Program provided with your membership, using a Performance Factor of 7 through 10, will work in perfect harmony with your maximum horsepower racing engine. An exhaust to intake port flow of 75% is ideal for racing engines that need to use the midrange torque band. Performance Factors of 4 through 6 can be used in the Header Design Program to maximize acceleration out of low RPM turns.

Cylinder heads for non-race applications often have exhaust port flow ratings of 68% to 73% of the intake port flow rating at maximum lift. These heads have more than adequate exhaust port flow for engine operation below the peak torque RPM, especially at part-throttle. But at full-throttle operation above the peak torque RPM, the horsepower potential of the engine might be limited by the exhaust port. The exhaust lobes on the camshaft must be properly selected to help compensate for poor exhaust port flow, although the peak exhaust port flow may still limit the maximum attainable power.

In some instances the widely accepted optimum port flow bias of 75% may not be the best ratio for your engine. Modern engines frequently have long intake runners designed to enhance midrange torque at the expense of high RPM horsepower. This type of engine can benefit from a flow bias of less than 75%, how much less depends on the application. Also, supercharged and restricted engines need a flow bias substantially higher and lower than 75% respectively.

The following guidelines should help you decide on valve sizes and porting effort when custom modifying a set of cylinder heads for naturally aspirated engines designed to make maximum horsepower. If your exhaust port flow percentage is below 70% of the intake, the engine will not respond to further improvements in intake port flow, but will show a drastic horsepower increase when the exhaust port or valve are improved. Conversely, if the exhaust port flow percentage is above 83%, the engine will not respond to further improvements in exhaust port flow, but will show a drastic horsepower increase when the intake port or valve are improved. If your exhaust port flow percentage is in the 71% to 75% range, the engine will be more responsive to exhaust improvements. If your exhaust port flow percentage is in the 78% to 82% range, the engine will be more responsive to intake improvements. If your fully prepared race heads have an exhaust to intake flow bias of 75% to 78%, then you can be confident that the engine will produce maximum horsepower. Always use valve and valveseat details that make the intake and exhaust ports flow poorly when tested for reverse flow if you want to improve engine performance below the peak torque RPM.

At low RPM the ram-charge effect is weak, and has little influence on the volumetric efficiency of the engine. The piston merely draws in fuel-air mix on the intake stroke, then pushes it back out of the cylinder on the compression stroke until the intake valve closes. But at the peak torque RPM, the cylinder is strongly ram-charged by this reflected compression wave after the piston has stalled near the bottom of its travel in the cylinder bore. Only a very small amount of the charge is pushed back out the intake valve on the compression stroke because of the presence of the ram-charging compression wave in the intake port. This allows compression to begin even though the intake valve is still wide open. At the peak horsepower RPM, the cylinder is ram-charged by the reflected compression wave when the piston is starting up the cylinder bore on the compression stroke. Essentially none of the intake charge is lost on the compression stroke at this high RPM if the camshaft has been correctly designed.

The intake runner cross-sectional area largely determines the ram-charge wave strength, and the runner length and taper angle largely determine wave timing. The desired ram-charge effect is greatly reduced if the intake runner cross-sectional area is too large, or has improper dimensions along its length. It is also reduced if the intake manifold volume is small (this doesn’t apply to individual runner applications). Runner size and length, and plenum volume must be carefully designed to provide an optimum combination of total manifold volume and wave strength. Since an engine operates over an RPM range, use an intake manifold that is designed to operate in that range. The intake manifold should have the same runner size as the cylinder head. Accurately match the port to the runner. If the intake manifold runners are larger than the ports in the head, the intake manifold was probably designed for a higher RPM range than the cylinder heads.

Although the design of runner lengths is always dominated by the timing of the ram-charging wave, residual waves continue to bounce back and forth in the runner after the induction period. If the end of a residual compression wave is present at the back of the intake valve the next time it opens (during the valve overlap period), the runner will be in a secondary state of tune that will complement the primary ram-tuning. If a scavenging wave is also present at the exhaust valve, the engine will be able to make maximum power at that RPM. The engine will go in and out of optimal intake tune as RPM changes.

The primary ram-charge wave timing dominates the intake runner design in maximum power engines that operate below 6500RPM. Residual wave tuning becomes important in race engines that operate above 6500 RPM. Lower speed street engines are frequently designed with long intake runners to produce maximum midrange torque. This reduces top-end power, and increases the importance of residual wave tuning. The intake runners must be carefully designed to produce the best torque curve for any given application.

Many types of intake manifolds are available, and all will work with headers. Intakes can be divided into two main categories: Individual runner (where the throttle plates are located in the intake runners), and plenum-type manifolds (where the throttle plates are remote to the intake runners). Various intakes are described below in order of decreasing ram-tuning potential:

Individual Runner Fuel Injection – Intake runner entries are directly exposed to the atmosphere, providing maximum ram-charge wave strength. No compromises need to be made when designing the optimum runner dimensions. Mid-range torque and top-end power potential are maximized. Throttle response is instantaneous.

Individual Runner Carburetion – Intake runner entries are directly exposed to the atmosphere, providing maximum ram-charge wave strength. Runner dimensions are slightly compromised to incorporate the carburetor venturi area. Mid-range torque potential is maximized, and top-end power potential is somewhat reduced because of the venturi restriction. Throttle response is instantaneous.

Port-Injected Plenum Manifold – Intake runner entries are located in a large plenum. With a total manifold volume that can approach 10 single-cylinder displacements in 8 cylinder engine applications, ram-charge wave strength can be near maximum. Ram-charging is reduced if runner entries are remote from the center of the manifold volume, when runner entry details do not allow peripheral feed, or when runners are aimed toward a plenum wall. No compromises usually need to be made, other than curvature, when designing the optimum runner dimensions. Mid-range torque and top-end power potential can be nearly maximized if the total manifold volume is large. Throttle response is slow because of the large manifold volume and remote throttle body.

Tunnel Ram Carbureted or Throttle-body Injected Manifold – Intake runners are positioned to enter the floor of a central plenum. Throttle bores are located directly above the runner entries. The location of the throttle bores strengthens the ram-charge, and improves fuel metering and distribution in carbureted applications. Total manifold volume can approach 10 single-cylinder displacements, providing near maximum ram-charge wave strength. No compromises need to be made when designing the optimum runner dimensions for this race-type manifold. Mid-range torque potential and top-end power are very nearly maximized. Throttle response is fast because of the location and number of throttle bores.

Single-Plane Carbureted or Throttle-body Injected Manifold – Intake runner entries are located in a small plenum where fuel is supplied. Total manifold volume is usually 5 to 7 single-cylinder displacements in V8 race engine applications. Ram-charge wave strength is strong, but less than that of the tunnel ram manifold. Correct fuel metering and distribution limits the total manifold volume potential of the single-plane manifold, especially in carbureted applications. The runner entries are centered in the manifold volume and are located very close to the throttle bores, both of which improve the ram-charge strength. Optimum runner dimensions may have to be somewhat compromised because of engine dimensions. Mid-range torque potential is modest to good, and power potential is good. Throttle response is generally good if the manifold is correct for the application.

Improved Dual-Plane Carbureted or Throttle-body Injected Manifold –Slightly oversized intake runners are divided into two groups, each group of runners sharing a common small plenum where fuel is supplied. The half-manifold volume is only about 2 to 2.5 single cylinder displacements, so powerful suction waves will reach the carburetor. Suction waves are partially dissipated by removing part of the divider that separates the two plenums, which improves the ram-charge strength. Oversizing the carburetor also improves ram-charging, but is generally an inferior solution to lowering the plenum divider. Ram-charge wave strength potential is modest. Optimum runner dimensions may have to be somewhat compromised because of engine dimensions. Mid-range torque potential is modest, and power potential is modest to good. Throttle response is generally good.

Standard Dual-Plane Carbureted or Throttle-body Injected Manifold – This manifold description also applies to many carbureted I-4 engines. Intake runners are divided into two groups on 6 and 8 cylinder engines, each group of runners shares a common small plenum where fuel is supplied. The half-manifold volume is only about 1.5 to 2 single cylinder displacements, so powerful suction waves will reach the carburetor at low engine speeds. Ram-charge wave strength potential is low in unmodified manifolds. Suction waves can be partially dissipated by removing part of the divider that separates the two plenums, which improves ram-charging. Throttle bore size also has a profound influence on ram-charging because of the small half-manifold volume. Optimum runner dimensions are usually compromised because of engine dimensions. Midrange torque potential is low, and power potential is modest. Throttle response and low RPM torque are generally excellent.

Obsolete Log-Type Carbureted Manifold – Intake runners are short, and plug into a log-type chamber mounted with one or more carburetors. This type of manifold was used before the 1950’s on common engines. The emphasis was on engine durability and accurate low-speed fuel metering, not horsepower. The ram-tuning potential of this type of manifold is minimal, but the cross-sectional area of the intake port is still very important. Midrange torque and horsepower potential are poor when compared to that of modern high-speed, ram-tuned engines.

A camshaft is often described as the brain of the engine. It controls the timing of valve open durations, as well as the rate the valves are opened and closed. These valve events have major influence on the performance of the engine. The importance of valve events is dominated by the timing of the intake valve closing, because this must match the performance level and RPM range dictated by the cylinder head. But the exhaust lobes on the camshaft must also be carefully designed in order to get maximum power and efficiency from the engine.

A correctly selected exhaust lobe will provide a smooth blow-down of cylinder pressure at high RPM that begins when the exhaust valve opens, and ends shortly before the piston begins to speed up the bore on the exhaust stroke. The camshaft designer must match the exhaust lobe opening rate and timing to the actual exhaust port flow measurements, to the cylinder head port flow bias, and to the overall design goal for the engine.

Essentially, some energy is stolen from the power stroke during blow-down, and used on the exhaust stroke to scavenge the cylinder. This trade-off is essential for making good horsepower in a performance engine. Engines with good exhaust port flow ratings will have slightly shorter exhaust valve open duration. Engines with poor exhaust port flow ratings will have slightly longer exhaust valve open duration, but not necessarily more lift. Maximum exhaust valve lift is partially determined by the required opening rate. Maximum exhaust valve lift must also be selected to provide consistent exhaust port velocity over the entire exhaust stroke at higher RPM levels.

The optimum combination of exhaust valve lift rate and maximum lift will allow the cylinder pressure to drop as steadily as possible on the exhaust stroke. Engines will generally benefit from a fast exhaust valve closing rate to maximize lift during the second half of the exhaust stroke. This allows the header to more thoroughly scavenge the cylinder at high RPM with an earlier exhaust valve closing time, and can provide a notable power increase if the valve springs can handle the closing rate. Our Header Design Program provided with your membership gives you the optimal header pipe diameter needed to complement an optimal cylinder blow-down. This allows the header to make a scavenging wave with the proper strength and timing for the engine. If your exhaust lobe releases an unnecessarily high amount of energy into the header, then the best header pipe diameter will be slightly larger than that calculated by the program.

The valve events will be discussed in order of decreasing importance:

Intake Valve Closing (IVC) timing – For a given well-designed engine combination, IVC time must be selected to provide quickest acceleration times over the design RPM range. IVC time has profound influence on the entire torque curve. As IVC time is delayed beyond bottom-dead-center, the engine’s torque curve will increase in magnitude, and will shift to an increasingly higher RPM range. There is a particular IVC time that will allow the engine to first reach its maximum horsepower potential. As IVC time is delayed beyond this, the horsepower curve for the engine will remain near maximum in magnitude, and will broaden to cover a wide RPM range. Peak torque will decline in magnitude. Also, there is a strong relationship between the ram-charge strength potential of the intake manifold and the optimal IVC timing. This is the primary reason why typical carbureted engines use a smaller intake lobe centerline angle than used in large plenum-type fuel injected engines.

Exhaust Valve Closing (EVC) timing – Engine power will be reduced if EVC timing doesn’t occur within a narrow range that is suitable for the engine. If EVC is too early exhaust scavenging will be incomplete at high RPM. This causes exhaust pinching during overlap, which drastically decreases horsepower. The engine will be able to make more power past the peak horsepower RPM as EVC is delayed, but EVC must still fall within the suitable EVC range. If EVC is delayed beyond the suitable range, the entire torque curve will decrease in magnitude. Street engines use earlier EVC timing to improve low RPM power at the expense of redline RPM power. Race engines use later EVC timing to maximize high RPM power near the shift RPM, thus extending the usable power band.

Exhaust Valve Opening (EVO) timing – Blowdown begins with EVO. The timing of this event largely depends on the flow bias of the cylinder heads, and can be used to move the engine’s torque curve up or down in RPM to a small degree depending on the design goals. EVO must fall within a range that is suitable for the engine. If EVO and the exhaust lobe are both correct for the application, the blowdown wave will peak in pressure in the exhaust port at about 10degrees before bottom-dead-center during middle to upper RPM engine operation.

Intake Valve Opening (IVO) timing – The intake valve is opened as early as possible while meeting the design goals for the engine. The earlier the intake valve starts to open, the larger the intake lobe will be on the camshaft. This is especially important for engines with weak ram-charge, since power output heavily depends on filling the cylinder during the intake stroke. Early IVO reduces the pressure differential across the intake valve during the first half of the intake stroke, so the intake runner is directly exposed to cylinder suction. The engine’s torque curve will change magnitude as IVO is changed, but the RPM range will not change. Early IVO also allows the header to more thoroughly scavenge the cylinder during valve overlap, and start induction while the piston is moving up the bore on the exhaust stroke. Later IVO must be used when exhaust port flow is poor, when part-throttle performance or fuel mileage is of prime concern, when correctly sized headers are not used, or at any time the exhaust system has the potential for backpressure at IVO.

When the piston compresses the fuel-air charge on the compression stroke, the piston does work on the charge. The charge’s pressure and temperature increase, and its overall energy increases. If the charge is not ignited, this negative work done by the piston will be recovered on the power stroke.

But if the charge is ignited, additional energy is added to the system. Pressure and temperature are increased. The higher the compression ratio, the higher the pressure will be at the beginning of the power stroke. This pressure does work on the piston as it descends down the bore on the power stroke. The temperature and pressure of the exhaust gas decrease as the piston descends, and the overall energy of the exhaust gas decreases. As compression ratio is increased in a given engine, the temperature of the exhaust gas blown into the header will decrease. This is because more of the total available energy is used for mechanical energy in the higher compression engine.

In real engines, compression ratio is limited by at least three concerns. The first is the mechanical integrity of the engine components. High compression engines generate much more force on the pistons than low compression engines, and therefore require stronger and more expensive engine components. The second limiting concern is combustion chamber shape. Very high compression engines are usually high RPM engines that require valve reliefs cut into the tops of the pistons. This keeps the pistons from impacting the valves during the overlap period. It also makes the combustion chamber very flat and irregular in shape. As a result, as compression ratio is increased the combustion chamber takes on an increasingly inferior shape for best ignition and combustion. A third limiting concern is detonation. Detonation can quickly destroy an engine if the compression ratio, volumetric efficiency, material and shape of the combustion chamber, and the octane rating of the fuel are not compatible. The probability of detonation must be evaluated at all RPM since volumetric efficiency, charge turbulence and flame speed change with engine speed.

Several factors serve to effectively change the engine compression ratio by changing the compression and combustion pressures. The first is volumetric efficiency. A second factor that effectively reduces compression ratio is the amount of residual exhaust gas in the cylinder and combustion chamber volume following the overlap period. Two cylinders having identical volumetric efficiency will produce different power depending on the amount of residual exhaust gas. This has a profound effect during part-throttle operation, and can greatly reduce fuel economy when dilution is high. A third factor is the pressure or air density available to the engine.

The Header Design Program uses your compression ratio input value when interpreting your peak horsepower. The program can then calculate the quantity of exhaust gas actually produced by the engine, and size the header accordingly. The thorough scavenging of residual exhaust gases provided by a properly sized header can give you a performance increase that feels like an increase in compression ratio.

The connecting rod length to stroke ratio has influence on the motion of the piston in the cylinder bore. In general, the piston tends to move the fastest about halfway down the bore, and quite slowly at top-dead-center and bottom-dead-center. If an exceptionally long connecting rod is used, the piston motion will be sinusoidal with respect to time or crank angle. The piston motion will be the same near top-dead-center and bottom-dead-center. Such an engine would have to be massive to accommodate the long connecting rod, and is therefore unrealistic. In order to reduce engine size and weight, the connecting rod length must be reduced.

As the connecting rod is shortened, the piston motion tends to “jab” at the top of the bore, and “stall” at the bottom of the bore. The side load on the piston from the rod also increases as the connecting rod is shortened, decreasing the life of the rings and bore and increasing internal engine friction losses.

For most engines a rod-to-stroke ratio of about 1.7 proves to be an optimal compromise between stall time near bottom-dead-center, which improves cylinder filling, and stall time near top-dead center, which improves combustion efficiency. Very high RPM engines often benefit from rod-to-stroke ratios higher than 1.7.

If you change your connection rod length, like when switching from 5.7inch rods to 6.0inch rods in a 350Chevy, your engine will require a different camshaft in order to maintain or improve performance. The longer rod starts the piston down-bore on the intake stroke later than the shorter rod. This retards the ram-charge effect and will reduce top-end horsepower. The intake valve closing time must also be retarded to compensate. Your exhaust lobe profile and centerline angle may also have to be changed. The header design produced by our Header Design Program will be accurate for any reasonable rod-to-stroke ratio as long as the correct camshaft is used.

The Header Design Program provided to our membership is only for naturally aspirated gasoline engines. The program assumes that the engine is built from a well-balanced set of components. The camshaft timing, and the cylinder head flow bias must be close to correct for the engine (near 75% exhaust port flow to intake port flow for maximum-output engines). Also, intake manifold pressure should be near atmospheric at high RPM.

When restricted carburetion or restrictor plates are used on race engines, the usual 75% flow bias built into most cylinder heads is no longer appropriate for the engine. A restricted engine is essentially a high-RPM, throttled engine. It requires a special camshaft to partially correct for the inappropriate flow bias built into the head. The header must also be specially designed to overcome the vacuum present in the intake manifold at high RPM, and help induction. Our Header Design Program has no way of knowing the degree of restriction, the flow bias, or if the camshaft has corrected for the inappropriate flow bias. Only a special design will produce a competitive restricted race engine. Our design service should be considered an essential element in matching your car and restricted engine to a particular track.

Supercharged engines have similar peculiarities to restricted engines. An average pressure other than atmospheric is present in the intake manifold at high RPM. Also, the flow bias built into the cylinder head might not be appropriate for the supercharged engine. The headers and the cam must be designed to compensate for, or partially compensate for, any weakness in exhaust port flow. Exhaust stroke pumping losses can seriously reduce supercharged engine output.

Header designs for turbocharged engines are even more complicated. Not only is the intake manifold pressure something other than atmospheric, the header outlet pressure is also higher than atmospheric on average. Furthermore, the flow bias in the cylinder head might be wrong for the engine. Residual exhaust energy and exhaust stroke pumping supply energy to the turbocharger, so exhaust port flow across the valve has an important impact on engine output, as does camshaft design and header size.

The intake signal at your carburetor will change for the better when properly designed headers are installed. This makes tuning the carburetor straightforward. Jet the carburetor a little rich until you have properly adjusted the primary and secondary throttle plates with respect to the transfer slots, and tuned the idle mixture and accelerator pumps. Start lean when tuning accelerator pumps and work toward rich. You may want to temporarily remove the power valve if equipped during the initial tuning of the accelerator pump, or when first firing a new carburetor/engine combination. Start conservative on secondary throttle opening time until you have tuned the primary side of the carburetor. A properly selected, sized, and tuned carburetor should work almost as well as fuel injection once the engine is warmed-up.

Read your spark plugs frequently while tuning your engine. Always read the ring of ceramic at the very deepest point of the plug (for conventional plugs). The rest of the ceramic won’t color for a long time. A light tan to light brown is about right for racing. Light gray or white is a little lean, even for economy applications, and will be down on power for heavy cruise or towing. Peppery specks on the end of the ceramic insulator near the electrode indicate a lean mixture condition. Fluffy black indicates a very rich mixture condition. If you have little bits of aluminum sticking to the electrode, you are detonating and need to retard the timing at some point, if the plugs otherwise read fine for color.

Since your engine will burn a more potent dose of fuel mix when proper headers are installed, you may have to retard the timing at part throttle, if not at full-throttle. Some exhaust in the fuel mix works as an anti-detonant, as does extra gas in the fuel mix. Check your spark plugs frequently for signs of detonation if you can’t hear your engine over an open exhaust system. Some fuel injection systems use knock sensors, and retard the timing automatically. A quality dynamometer tune-up is your best bet for top performance.

The scavenging wave in a primary header pipe can have appreciable intensity, even under part-throttle operating conditions. In a properly designed header, a scavenging wave will be present at the exhaust valve before and during the valve overlap period under part-throttle operating conditions at cruise RPM. This inhibits flow of exhaust gases into the intake manifold, where vacuum is high at part-throttle. As a result the intake charge will have less contamination from heat-stealing exhaust gases. This results in an increase in power output per amount of gasoline, and is similar to raising compression ratio. Throttle response will be sharper at cruise, and intake manifold vacuum will increase.

Engines equipped with exhaust manifolds or shorty headers cannot provide intense scavenging of the combustion chamber during valve overlap. They tend to dampen exhaust pressure waves. As a result the intake charge is always contaminated with the exhaust gases that remain in the combustion chamber. These exhaust gases are at a higher pressure than that in the intake manifold under part-throttle conditions, and tend to enter the intake port. If a performance camshaft is used with long valve overlap duration, a significant amount of exhaust gas can be drawn into the intake manifold out of the exhaust system, ruining part-throttle economy.

Under low RPM light-throttle conditions, a long-tube header pipe limits the quantity of exhaust gas that can be drawn into the intake manifold. During valve overlap a strong suction wave is initiated from the intake port, travels across the combustion chamber, and out the exhaust port. Even at low RPM time does not allow a compression wave, reflected from the header collector, to arrive back at the exhaust valve before the valve overlap period ends. This results in minimal exhaust contamination of the intake charge during overlap (unless the header is oversized), and makes carburetors much more responsive to your light-throttle tuning efforts. When exhaust manifolds or shorty headers are used, time does allow for a compression wave (reflected from the nearby tailpipe or common chamber) to arrive back at the exhaust valve during valve overlap. The result is more exhaust contamination during overlap, especially with performance overlap durations, and lower intake manifold vacuum under low RPM light-throttle conditions and at idle.

Your header design must be matched to your cruise RPM if economy is important. If you use a header that is too short for your application, and especially if the collector is too large, you may find that your economy will decrease. This is because the scavenging wave created in the header is followed by a compression wave. This compression wave will arrive during valve overlap at cruise if the header pipes are too short, and inject exhaust back into the cylinder and up the intake runners. Soggy throttle response and poor power at cruise will be the result. You may find evidence of exhaust gas contamination on the floor of the intake manifold runners if you consistently cruise with an out-of-tune header design. Our Header Design Program produces header dimensions that will dampen this compression wave when it is present during overlap, which improves low-RPM performance.

Headers designed with the Header Design Program using a Performance Factor of 1 will have primary header pipe lengths long enough to provide good scavenging at a cruise RPM that is a minimum of about 55% of your peak horsepower RPM. A Performance Factor of 4 will require a cruise RPM of about 65% of your peak horsepower RPM. For example, a truck with 27” tires, 3.08gear, and automatic transmission will cruise at about 2600RPM at 65mph. If the peak horsepower RPM is 4500, then cruise RPM is 57% of that. Use a Performance Factor of 2 for best economy.

If you use an overdrive gear to cruise, the header needs to be specially designed to provide the second returning scavenging wave during overlap at your cruise RPM. The overdrive gearing will give best economy when very little engine power is required to cruise. This is because engine friction horsepower losses are minimized at low RPM. If you have a high power requirement at cruise, as is the case for tow vehicles with high wind drag, you will need to cruise near the peak torque RPM for best economy. Once the cruise RPM has been determined and the headers designed, part-throttle fuel efficiency will be maximized by using the highest compression ratio that will work with your fuel, and by minimizing valve overlap duration (use later intake valve opening, and optimal exhaust valve closing).

Stepped-tube headers are also useful when a race engine’s operating range needs to be extended well above the peak horsepower RPM. Engines for stock cars raced on low-banked intermediate tracks, and engines used for all-out drag racing might fall into this category. HeaderDesign.com subscribers will benefit greatly from the instructions on when to use, and when not to use, stepped-tube headers.