[ -> siehe auch diesen Beitrag ]

„Der Auspuff, mysthisches heißes Rohr. Oh, warum bist du nur so laut? Wo bleibt der erwartete Leistungszuwachs, alte Knalltüte!?“

Ganz klarer Fall von KEINE AHNUNG. Damit das besser wird und nicht mehr so weh tut, habe ich mir einige Infos zusammengesucht, wie man einen Auspuff, besser gesagt den Krümmer, auslegt. Mit Laienmitteln. Immerhin.

Bei allen Rechnungen sollte übrigens auf die Anzahl der Zylinder aufgepasst werden, auf die sie zugeschnitten ist. Und: Es sind immer nur Pi-mal-Daumen-Werte die herauskommen, auch wenn die Berechnungen noch so hochtrabend erscheinen mögen.


„Tuned exhaust  (From Wikipedia, the free encyclopedia)

A tuned exhaust system is an exhaust system for an internal combustion engine which improves its efficiency by using precise geometry to reflect the pressure waves from the exhaust valve or port back to the valve or port at a particular time in the cycle.

Tuned exhaust systems are common on motor racing engines, light aircraft engines, model engines and two-stroke motorcycle engines, and are not restricted to these applications. They are of two main types:

* In a two-stroke engine, tuned expansion chambers are used to reduce loss of the new charge caused by late closing of the exhaust port by delivering a pulse of positive pressure after the exhaust gases have left the cylinder.
* In a four-stroke engine, tuned extractor manifolds are used to promote scavenging of the exhaust gases by delivering a pulse of negative pressure just before the exhaust valve closes.

In either case, the efficiency of the system is greatest at a particular engine speed, which is determined by the exhaust system geometry. Tuning the system for best effect is typically done both by calculation during the design of the exhaust system, and by trial and error during engine development.



Berechnungen – Link 1         http://www.mezporting.com/exhaust_length.html

Dabei im Hinterkopf behalten:

„(…)  Since all numbers are approximate, let me explain to you the concepts you have to understand behind the design. The pressure wave begins its journey at the exhaust valve. It then travels at the speed of sound throughout the pipe. The speed of sound varies with exhaust gas temperature, so the pressure wave will be slower at the end of the pipe than at the exhaust valve (It is even more complex than that, but what you have to remember is that the tuned RPM is not directly related to the pipe length; other factors are involved). Once the pressure wave encounters an area change, it will reflect another wave back to the engine. If the area is reduced, it will reflect a pressure wave slightly smaller (if it hits a closed end, it will bounce of the end and be reflected as the full pressure wave). We don’t want that. If the pressure wave encounters a larger area, it will reflect a «suction» wave: The bigger the pressure wave and the bigger the area change, the bigger the «suction» wave. Once this «suction» wave arrives back at the engine, it will help emptying the cylinder. The largest area change possible is at the end of the pipe, where it meets the atmosphere (area is infinite at this point). If you have a pipe going into a bigger pipe, the pressure wave will partly reflect as a «suction» wave and it will also continue its journey into the bigger pipe, slightly diminished. This slightly diminished pressure wave will also reflect another «suction» wave when it encounters another area change along the way in the bigger pipe. Knowing all of that, here’s how you have to analyze your design: If you have an exhaust pipe of a single diameter – equal to the exhaust port diameter – and this pipe discharges directly into the atmosphere, you will have a very peaky engine, tuned for the RPM you’ve calculated. The drawback is that it will perform poorly at other RPM. This is how it is done for racing engine looking for maximum horsepower (like a dragster for example). If you want a larger power band, you need to have «different» pipe lengths for the RPM range you desire. So you can increase the pipe diameter along the way. Ultimately, you can use a tapered pipe, which will reflects «suction» waves at every millimeter of your pipe. Naturally, since the area change is small at every step, the «suction» wave reflected is also small and offers less maximum power at a given RPM that a straight pipe discharging into the atmosphere. How big of an area change do you need ? To «simulate» an atmosphere, you need a diameter change of 2,5 or bigger. So if you have a pipe of 1″ in diameter discharging into a pipe of 2,5″ in diameter, for the pressure wave it will feel like discharging into the atmosphere; meaning that any other area change afterward will be more or less meaningless as the pressure wave will practically completely loose its strength at this point.“



Berechnungen – Link 2, Software zum Kaufen.  Soll die absolute Geldverschwendung sein…



„Header Basics

Think you know all there is to know about headers? Grab your favorite brain food and get ready to pay attention. We’ll discuss all aspects of header design so you can accurately choose which set of pipes is best for your motor to exhale through.

Frankly, you’d think this subject would have been exhausted by now. After all, how much “borderless education” can you absorb about such common and oft-explained engine functions as getting rid of combustion by-products? Well, this story offers you a challenge. Our plan is to integrate various header functions, dispel a few myths about how headers work, and simplify matching parts to engine size and rpm.


Many stock exhaust systems are not capable of transferring sufficient exhaust gas at high engine speeds. Restrictions to this flow can include exhaust manifolds, catalytic converters, mufflers, and all connecting pipes routing combustion residue away from the engine. As power levels increase, proportionate amounts of exhaust can also increase, placing added demands on systems that may be flow deficient. Header manufacturers, among other objectives, attempt to build systems that fit (or should) and provide bigger pipes for high-rpm power gains. Knowing how and why a system needs to work helps in the selection process.

Combustion by-products won’t burn a second time. Therefore, an exhaust system that cannot properly rid cylinders of exhaust gas can cause contamination of fresh air/fuel charges. Residual exhaust material occupies space in the cylinders that prevents maximum filling during inlet cycles. As a rule, this problem grows with rpm, potentially reducing the benefits that can be derived from other performance-enhancing parts.As you will see, exhaust-flow velocity is an important component in an efficient exhaust system. Simply stated, at low rpm, the flow rate tends to be slow.

As engine speed increases, so does flow rate. Then, as restrictions increase, velocity slows again, reducing power accordingly. Interestingly, camshaft design, compression ratio, ignition-spark timing, and piston displacement affect all this if an accompanying improvement in the exhaust system isn’t included with such changes. In fact, these types of modifications can cause exhaust problems to occur sooner in the rpm range.On the other hand, exhaust systems can be too big for engine packages that don’t produce sufficient exhaust-flow volume to necessitate size increase. So we’re back to the flow-velocity issue. Sizing of system components, such as headers, can be keyed to engine speed and piston displacement. We’ll show you how this is done later in the story.

Graph A illustrates how merely changing pipe diameter affects an engine’s output. Note that the smallest diameter creates good midrange torque yet falls off at the top, while the larger primary header pipes add more high-rpm power at the expense of low-speed torque.

Primary pipe length can also skew an engine’s power curve based on length changes. Primary-pipe diameter establishes the peak torque point, so changing the pipe length will rock the output curve by pivoting it around that peak torque point. Graph B shows how longer tubes tend to increase power below peak torque while hurting power above peak torque. Shorter tubes tend to affect the engine in exactly the opposite way, hurting midrange torque in favor of increasing top-end power.

What Primary Pipes Do

The main function of primary pipes is to set the initial rpm point (engine speed) at which a torque boost is created, as contributed by the headers. Keep in mind, exhaust and intake systems can be tuned to different engine speeds. By so doing, an overall torque curve can be broadened or narrowed by the separate dimensioning of intake and exhaust systems.

Several variables contribute to how headers affect engine performance, including the diameter and length of the primary-pipe and collector.

In the case of headers, primary-pipe diameter determines flow rate (velocity). At peak torque (peak volumetric efficiency), the mean flow velocity is 240-260 feet per second (fps), depending upon which mathematical basis is used to do the calculation. But for sizing or matching primary pipes to specific engine sizes and rpm, 240 fps is a good number.

Changing the length of primary pipes generally affects the amount of torque produced above and below peak-torque rpm. For example, all else being equal, shortening primary pipes transfers torque from below to above the peak, not significantly shifting the rpm point at which peak torque occurs. Increasing primary-pipe length produces the opposite effect of shortening the length.

Primary-pipe diameter plays a big part in determining the torque curve. A pipe that is too small creates peak exhaust-gas velocity too early in the rpm curve and will limit top-end horsepower. Pipes that are too large create a very peaky output curve.

What Do Collectors Do?

Essentially, collectors have an impact on torque below peak torque. While the gathering or merging of primary pipes does affect header tuning, it is the addition of collector volume (typically changes to pipe length once a diameter is chosen) that alters torque. Engines operated above peak torque, particularly in drag racing, do not derive any benefit from collectors. Those required to make power in a range that includes rpm below peak torque do benefit. And the further below peak torque they are required to run (from 2,500-7,500 rpm for example), the more improvement collectors provide.

Joining collectors, cross-pipe science notwithstanding, tends to further boost low-rpm torque by the increase in total collector volume. Generally, crossover pipes become less effective at higher rpm, as you might expect, although some manufacturers of the more scientific cross-pipes claim power gains as engine speed increases. The mere joining of collectors in a dual-collector system does not appear to produce this improvement.

Header Size

Consider this: It is the downward motion of a piston that creates cylinder pressure less than atmospheric. Intake flow velocity then becomes a function of piston displacement, engine speed, and the cross-section area of the inlet path. On the exhaust side, a similar set of conditions exists. In this case, exhaust-flow velocity depends on piston displacement, engine speed, the cross-sectional area of the exhaust path, and cylinder pressure during the exhaust cycle.

Of the similarities between the intake and exhaust process, piston displacement, engine speed, and flow-path cross section are common. Therefore, there must be a functional relationship among rpm, piston displacement, and flow-path section area, and there is (see the section on calculating pipe sizes).

Note how this shorty header minimizes the collector length. Generally, this is done to make it easy to fit the headers in the chassis. This mini-collector generally costs low-speed torque. If you have room to extend the collector length by 8 to 10 inches, this could improve torque.

This suggests the possibility of sizing primary-pipe diameter to produce torque boosts (as contributed by the exhaust system) to an engine’s net torque curve. The previously mentioned mean flow velocity (240-260 fps) found in primary pipes around peak torque rpm is a function of pipe diameter. So, selecting diameters that correspond with the rpm at which torque boosts are desired is one method of header selection or sizing.

Matching Headers to Objectives

If we know any two of the three previously mentioned variables (piston displacement, rpm, or primary-pipe diameter), we can apply some simple math to solve for the other. Here’s how that works.

1. Peak torque rpm = Primary pipe area x 88,200 / displacement of one cylinder. Given this relationship, we can perform some transposition to solve for the primary-pipe cross-section area.

2. Primary pipe area = peak-torque rpm / 88,200 x displacement of one cylinder. We can also determine the required displacement of one cylinder (multiplied by the number of cylinders for total engine size) by:

Out of all the variables to consider, one of the most important is that the headers fit the chassis without requiring significant surgery. There are also several different variations on the angle-plug concept for small-block Chevy engines that can cause difficulties if you haven't done your homework.

3. Displacement of one cylinder = Primary pipe area x 88,200 / peak-torque rpm.

Equations 1 and 2 provide a method for determining peak-torque rpm (as contributed by the primary pipes) if you have already selected a set of headers and know the engine size. In equation 3, primary-pipe area can be determined if the desired peak-torque rpm and engine size are already known. It will also calculate engine size based on a known set of headers and rpm at which peak torque is desired.

Here’s an example of how this approach can work. Suppose you have a 350ci small-block (43.75 cubic inches per cylinder). A primary-pipe torque boost around 4,000 rpm is your target engine speed. The choices for pipe size are 15⁄8 inches, 13⁄4 inches, and 17⁄8 inches. If we assume a tubing wall thickness of 0.040 inch, each of these od dimensions requires subtracting 0.080 inch when computing cross-section areas.

Using the formula, Area = (3.1416) x (id radius) x (id radius), we obtain the following cross sections: 15⁄8 inches = 2.07 square inches; 13⁄4 inches = 2.19 square inches; 17⁄8 inches = 2.53 square inches.

Remember that headers are just one part of the power equation. When trying to improve power, consider the headers as one part of the overall compression, cylinder-head, camshaft, and induction system.

Plugging each of these values into equation 1, we find the selection of peak torque becomes (in the same order of pipe sizes), 4,173, 4,415 and 5,100 rpm. Based on an intention to provide a torque boost around 4,000 rpm, 15⁄8-inch-diameter primaries appears to work. In accord with our previous comments about primary-pipe length, extending these primaries will increase torque below 4,000 rpm at the expense of torque above this point, which is an additional tool to manipulate a torque curve about its peak (see “Torque Peaks”).

While this method will not predict header-pipe area as precisely as some contemporary computer-modeling programs, it can be a valuable quick-and-dirty tool when making decisions about header choice or application of sets already on hand.


There is much more to the science of exhaust-system tuning and headers that space does not allow us to include. It’s worth noting once again that the final combination of parts must take into account all the components as a system, rather than looking at the headers as a separate entity. Any engine will make its best overall power when treated as a complete system.



– Exhaust System Technology-

The Sound and The Fury



RACE ENGINE TECHNOLOGY MAGAZINEAll too often the engine exhaust is an afterthought for the engine and chassis builders, yet its design and construction impacts significantly upon car performance. The exhaust system can be a vital tool for optimizing the performance of the engine, through the way in which its design manipulates the pressure waves that can crucially assist cylinder filling and scavenging. On the other side of the coin, the exhaust system presents many challenges. It is a major loss-path for thermal energy; and it can be a car packaging nightmare.

The environment which a competition exhaust system, and particularly engine headers, must survive, can only be described as a brutal combination of temperatures, stresses, corrosion and vibration. Contemporary exhaust technology can help reduce the problems and help to maximize the potential gains of the system.

BMW Formula-One Engine at Full Power

Figure 1
BMW Formula-One Engine at Full Power

It is interesting, from having spoken to several highly-placed and well recognized experts in this field, that while there is general agreement about what features cause improvements to happen, there are varying opinions about the reasons why those improvements occur.


The computation of what actually goes on during an exhaust cycle is a highly complex problem in compressible fluid flow, the details of which are explained in detail in several texts, my favorite being Professor Gordon Blair’s Design and Simulation of Four Stroke Engines. For the purposes of this article, the following overly-simplified explanation will serve to illustrate the principles.

There are two separate components to the exhaust event. The first is the removal of exhaust gasses from the cylinder, which occurs as a pulse of hot gas exiting the cylinder and flowing down the header primary tube. The second is the (much faster) travel of the pressure wave in the port caused by the pressure spike which occurs when the exhaust valve opens, and the various reflections of that wave. Taking proper advantage of these pressure waves (component two) can produce dramatic improvements in clearing the cylinder (component one) and can strongly assist the inflow of fresh charge.

Considering component one, when the exhaust valve first opens in a 4-stroke piston engine, the in-cylinder pressure is still well above atmospheric. In a normally-aspirated spark ignition engine burning gasoline and operating at high BMEP, the pressure can be 7 bar or more, and the pressure in the exhaust port at the valve is somewhere near 1 bar (atmospheric). As the valve opens, the pressure differential across the rapidly-changing valve aperture (pressure ratio of approximately 7) starts exhaust gas flowing through the opening, and the outrush causes the pressure in the port (behind the valve) to increase rapidly, or „spike“.

The instantaneous velocity of the exhaust gas flow at any point is determined by the pressure gradient and the cross-sectional area at that point. In the header, a smaller tube diameter will increase the velocity at a given RPM, which might enhance the pressure wave tuning (the second component) and can be beneficial with regard to inertia effects. However, if the diameter is too small, there will be flow losses and consequent pressure gradient increases which can offset any tuning gains. So the selection of proper tubing diameters is an important part of the design.

In the early part of the exhaust cycle, the pressure difference across the valve is high, so the instantaneous gas particle velocity through the small exhaust valve aperture is very high. Sometime past mid-exhaust stroke, the majority of the exhaust gas has left the cylinder. At that time, the valve aperture area is quite large and the cylinder pressure is approaching atmospheric, which causes the instantaneous particle velocity across the valve to be much lower. It is at that phase of the exhaust cycle where the second component becomes important.

To help with the explanation of the second component, Figure Two shows traces of in-cylinder pressure (black), port pressure at the intake valve (light blue) and port pressure at the exhaust valve (red), taken from a simulation of a high BMEP engine operating near the optimum tuning point for both intake and exhaust.

Intake Port, Exhaust Port and In-cylinder Pressures with Effective Tuning

Figure 2
Intake Port, Exhaust Port and In-cylinder Pressures with Effective Tuning

The second component is the result of the pressure „spike“ which occurs at EVO, shown by the peak in the red line in Figure Two, just after EVO. That pressure spike, or pressure wave, moves down the pipe at the sum of the local sonic velocity plus the particle velocity of the gas flow. Whenever the pressure wave encounters a change in cross-sectional area of the pipe, a reflected pressure wave is generated, which travels in the opposite direction. If the change in area is increasing (a step, collector, the atmosphere), the sense of the reflected pressure wave (compression or expansion) is inverted. If the change in area is decreasing (the end of another port having a closed valve, or a turbocharger nozzle, for example), the sense of the reflected wave is not inverted. The amplitude of the reflected wave is primarily determined by the proportionate change in cross-sectional area (area ratio), but the amplitude is diminished in any case. For purposes of approximation, the particle velocity can be ignored because its effect is self-canceling during the round-trip of the wave. However, highly-accurate simulations must take it into account. These waves are sometimes called finite difference waves, because of the finite difference numerical modeling techniques used to calculate their propagation characteristics.

In the case of the currently-flowing header primary, the EVO-initiated positive pressure (compression) wave is reflected back as a negative pressure (expansion) wave. If the arrival of the reflected negative pressure wave back at the exhaust valve can be arranged to occur during the latter part of the exhaust cycle, the resulting lower pressure in the port will enhance the removal of exhaust gas from the cylinder, and will reduce the pressure in the cylinder so that when the intake valve opens, the low pressure in the cylinder begins moving fresh charge into the cylinder while the piston is slowing to a stop at TDC.

Note in Figure Two, how the cylinder pressure (black) and exhaust port (red) pressures go strongly negative from approximately mid-exhaust stroke to TDC). Note also how the second-order reflected positive pressure wave in the intake tract (light blue) reaches the back of the intake valve just before IVO, and works together with properly-timed exhaust negative pressures to begin moving fresh charge into the cylinder.

If, on the other hand, the negative exhaust pressure wave arrives a non-optimal time, its effects can be detrimental to the clearing of the cylinder and ingestion of fresh charge. A reflected positive wave during overlap (from a turbocharger nozzle, for example) can push a large amount of exhaust gas back into the cylinder and the intake system.

Figure Three shows the same three pressure traces when the engine is operating well above the intake and exhaust tuning points. In addition to reduced breathing efficiency, note the additional pumping losses from the higher cylinder pressure in the latter portion of the exhaust cycle, caused in part by the late arrival of the reflected negative exhaust pulse.

Intake Port, Exhaust Port and In-cylinder Pressures with Poor Tuning

Figure 3
Intake Port, Exhaust Port and In-cylinder Pressures with Poor Tuning

The timing of the arrival of the negative wave at the back (port) side of the exhaust valve is determined by the engine RPM, the speed of sound in the pipe and the distance from the valve to the relevant change in area. Those three factors will cause the exhaust tuning to come in and out of tune over the engine operating speed range. Sophisticated designs can produce systems having more than one tuning point. The most significant example of exhaust pulse tuning is dramatically demonstrated by the operation of crankcase-scavenged, piston-ported two-stroke engines.

At the relevant tuning distance from the exhaust valves, the primary tubes from two or more cylinders are often joined together into a larger collector tube which provides the area increase to generate the reflected waves described above.

Using a 4-into-1 system as an example, the four primary tubes will ideally have the same centerline length and will sharply transition into an area having roughly three to four times the area of the primary.. The larger the cross-sectional area of the collector tube plus the area of all other tubes at the same junction compared to the area of the active primary tube (area ratio), the larger will be the amplitude of the reflected wave. However, the collector has an optimal size: too much area and the wave tuning in the collector will be diminished. The optimal length is related to the number of cylinders feeding into it.

The effect of a straight collector is generally a very peaky tuning, in which the lengths of the primaries can be varied to produce a ‚rocking‘ effect of the torque curve around its peak. Lengthening the tubes raises the portion of the curve below peak and reduces the portion above peak torque; shortening them has the reverse effect. Various strategies have been devised to spread the effect of exhaust tune over a wider RPM band. These strategies typically involve generating additional waves of smaller amplitude (additional, smaller steps, for example) or attempts to increase the width (duration) of the pulse at the expense of pulse amplitude by using a tapered section to extend the area change over a longer period of time.

Figure Four shows one of these devices, known in the States as a ‚merge collector‘. The primaries converge into a nozzle area which is larger than the primary area but smaller than the final collector size. That keeps the gas velocity up for a bit longer, helping to scavenge neighboring pipes, and the smaller area ratio reduces the amplitude of the reflected wave. The section behind the nozzle tapers up to the final collector diameter, allowing the flow to decelerate with better pressure recovery than would occur with a sharp transition, and extends the width of the reflected wave. The characteristics of the reflected wave can be tuned with different nozzle areas, different final collector diameter and length, and the length of the tapered section. The net effect is usually aimed at boosting a particular portion of the torque curve and at extending the RPM band over which that boost is effective.

A "Merge Collector"

Figure 4
A „Merge Collector“

It is sometimes argued that the speed of sound is a function of pressure, density, temperature, and / or phase of the moon. Actually, the speed of sound in an ideal gas (which air emulates) is a function of the stiffness of the gas divided by the density. When one does the arithmetic necessary to create an equation which uses known parameters, the stiffness and density terms are replaced by equivalents from the ideal gas law, producing the equation: Va (acoustic velocity in meters per second) = square root ( S x R x T), where S is the ratio of specific heats (approximately 1.4 for air at 25°C, 1.35 for exhaust gas at 500°K), R is the gas constant (approximately 287 J/kg-°K for air, 291 for exhaust gas) and T is the absolute temperature (°Kelvin, which is °C + 273).

What that boils down to is that once one has the specific heats and gas constant value for a given gas (or mixture of gasses), the speed of sound varies only with the temperature. To add a bit of complexity, the instantaneous temperature of the exhaust gas varies along the exhaust path, perhaps as much as 150°C in a primary tube.

The next interesting basic is that as the pressure ratio increases across a smoothly-decreasing nozzle, the particle velocity at the smallest cross-sectional area increases with increasing pressure ratio until it reaches the local speed of sound. Once it has reached the speed of sound, no matter how much larger the pressure ratio becomes, the gas particle velocity remains at sonic („choked“). An increase in the upstream pressure will increase the mass flow rate due to the increased density upstream of the nozzle, but the particle velocity through the nozzle remains sonic.

For air flowing in a smoothly-decreasing nozzle, the pressure ratio which just causes sonic flow (the ‚critical pressure ratio‘) is slightly less than 2.0. For non-smooth and irregular nozzles (an exhaust valve, for instance) the critical pressure ratio is higher, but the effect is the same. That means that, for some period of time after EVO, the gas particle flow velocity across the exhaust valve is at the local speed of sound, which as shown later, is quite high at exhaust gas temperatures.

Again, it should be noted that these explanations are highly simplified. There are several very-high-end engine simulation software packages which are said to model engine performance, including exhaust system phenomena, quite accurately. These models are so sophisticated that they can take into account such esoterica as the local temperature gradients along the primary, secondary and collector tubes. For accuracy, these models rely on accurate engine data, including valve flow coefficients at various lifts. Apparently it is difficult to determine accurate flow coefficient data for the valves, particularly at high pressure ratios, which has a profound influence on the limits of computational accuracy.

That being said, several designers told me that the simulations tend to be less accurate in predicting the various effects of the collector, in terms of the real world effects of geometry, pipe angles, and the like. One approach to that problem has been to use a CFD simulation (a 3-D analysis) for the collectors, and couple those results with the 1-D simulations of the pipes.

Exhaust Materials

Usually, header systems are fabricated from welded-up collections of cuts from pre-formed „U“ bends and straight segments of tubing in the chosen material. There are several reasons for that, but the most persuasive is the fact that, in order to achieve the design configuration, there is not usually ample grip-space between bends to form the pipes from a single piece of tube. In some cases, where the bends are not too closely spaced, the pipes can be bent up in one piece using a mandrel bender which will retain the circular cross section of the tube throughout the bend and transition. The typical exhaust tube bender commonly found in automotive exhaust shops is not suitable for that duty since those benders distort the cross-section of the bends terribly and shrink the cross-sectional area.

Tubing bend radii (the radius of the plan-view centerline of the bend) are expressed in terms of multiples of the tubing diameter. For example, a „1.5-D bend“ in 2-inch diameter tubing would have a bend radius of 3 inches. One fabricator described some specialized machinery he had devised for making high-quality exhaust tubing from sheet. The first machine rolls the sheets into straight tube sections of the required diameter. The second machine completes the straight section of tube with a continuous welded seam using a semiautomatic inert-gas-shielded process. A third machine does what had been thought to be impossible: bending 0.50-mm wall inconel tubes into less-than-1-D radius sections while retaining accurate cross-sectional geometry.

There are several materials commonly used in competition header and exhaust systems, depending on the requirements and operating temperatures.

For the most demanding applications, Inconel tubing is commonly used. Although the name „Inconel“ is a registered trademark of Special Metals Corp., the term has become something of a generic reference to a family of austenitic nickel-chromium-based superalloys which have good strength at extreme temperatures and are resistant to oxidization and corrosion. Because of the excellent high-temperature properties, Inconel can offer increased reliability in header systems, and in certain applications, it is the only material which will do. The high-temperature strength properties can enable weight-reducing designs, since, for a given reliability requirement, Inconel allows the use of much thinner-wall tubing than could be used with other materials. The catch, as usual, is that Inconel tubing is quite expensive.

Certain Inconel alloys retain very high strength at elevated temperatures. One of the favorites for header applications is Inconel-625, a solid-solution alloy containing 58% Nickel, 22% Chromium, 9% Molybdenum, 5% Iron, 3.5% Niobium, 1% Cobalt. It has good weldability using inert-gas-shielded-arc processes, and good formability in the annealed condition, and has a lower thermal expansion rate than the stainless alloys commonly used in exhaust systems. Weldability and formability are both important because of the somewhat limited availability of Inconel tubing sizes, which often makes it necessary to form tubing sections from sheet. The yield strength of this alloy at 650 °C (1200°F) is 345 MPa (50 ksi), while at 870°C (1600°F) it is a remarkable 276 MPa (40 ksi). As with many metals, the high-temperature strength diminishes as the amount of time the parts are exposed to extreme temperatures increases.

Inconel tubing is nearly essential in high-output turbocharged applications, and I was told by several knowledgeable players that all the Formula-One cars and a few Cup teams use Inconel for their headers, both for reliability and for weight savings.

One builder told me that some teams are routinely using headers made from 0.50-mm (0.020 inch) wall Inconel tubing. He also told me that, in view of the immense heat load imposed by the exhaust gasses of contemporary Formula One engines, he seriously doubted that a set of stainless headers, even in 1.6-mm wall (0.065 inch), would survive. Figure One, a BMW F-1 engine at full power, graphically illustrates this demanding environment.

There are several austenitic stainless alloys which are commonly used in exhaust systems. In order of reducing temperature capabilities, they are 347, 321, 316 and 304. In addition, special variations in the basic alloy chemistry (carbon, nickel, titanium and niobium) are available to enhance the high temperature strength of these alloys.

Regarding the use of stainless, I was told by a knowledgeable source that in NASCAR Cup racing, the 304 and 321 stainless alloys were used more often than Inconel, depending on the preferences of the various teams. The manager of one prominent team told me that, in view of the facts that thinwall Inconel headers are (a) very fragile and readily damaged by inadvertent mishandling, (b) „grotesquely“ expensive, and (c) provide almost immeasurable gains on a 3600 pound vehicle, his opinion is that the use of Inconel headers is not prudent stewardship of his resources. For a peek at the magnitude of the costs involved, one fabricator told that a single 1-D „U“ bend of 2-inch diameter, 0.032-inch wall Inconel tubing would cost somewhere in the neighborhood of $200, whereas the same bend in 321-stainless would be in the $65 range.

Although titanium has been made to work quite well in exhaust valve applications, the practical temperature limits for titanium alloys suitable for tubing is quoted at about 300 °C (575 °F), which makes that material suitable for lightweight tailpipes in various applications and in certain motorcycle applications as well. My favorite supplier of titanium reports that grades 1 and 2 commercially-pure (CP) titanium have been used for the exhaust systems on competition 2-stroke motorcycles for decades. For lightness, many of these systems were made using 0.50 mm wall tubing, and treated as a consumable, being replaced after every meeting.

One might ponder why the same materials used for titanium exhaust valves are not used for exhaust tubing. Apparently, the simple reason is cost vs. benefit, since the estimated cost of thin sheets of Ti-6242 were estimated at over $150 per pound in large-quantity purchases. Add to that the fact that this material lacks the ductility to be readily formed into tubes, plus the fact that there would be problems welding the seams of a rolled tube, and yet more problems forming the welded straight tubes into bends, and it becomes evident that there are more suitable materials for exhaust tubing use.

Formula One

Recently, I had the opportunity to hold in my tired, worn hands, a primary header tube which was alleged to have been for a nearly-contemporary F-1 application. Pictures of said hardware were not allowed, but the reproduction from memory, shown in Figure Five, illustrates the very interesting feature, the existence of a large-diameter step in the primary, quite close to the flange.

Formula-One Primary Header Tube

Figure 5
Formula-One Primary Header Tube

The illustration shows a single 10-mm step spaced approximately 125 mm from the flange. However, experts say that in 2008, two smaller steps (5 mm each) in the primary are more commonly seen, depending on the research and beliefs of the developers. The first step is typically between 100 and 200 mm from the flange. If there is a second step, it is typically another 100 to 150 mm beyond the first step, and in general, tubing sizes range from about 50 mm to 65 mm. (1.97″ to 2.56″), although the specific designs seem to vary dramatically from team to team.

My first impression, which was shared by a number of experts with whom I spoke, was that, since these engines are operating up to 19,000 RPM, then the primary length required to achieve the negative pressure pulse during overlap was so short that, due to packaging constraints, the location of the collector would be too far away from the valves to initiate the properly-timed reflection. However, a bit more thought and a quick calculation revealed quite a different theory.

For purposes of approximation, assume that the mean temperature of the exhaust gas in the primary up near the head is 1500°F (815 °C). The speed-of-sound-in-air equation (close enough for approximations, according to Professor Blair) produces a sonic velocity of 661 m/s (2168 feet per second). At 18,000 RPM, (300 RPS) one crankshaft rotation takes 3.33 milliseconds (ms) or 3333 microseconds (μs). Therefore one degree of crank rotation takes 9.26 μs (3333 ÷ 360). If the first step in the primary is 200 mm from the back of the exhaust valves, then using the calculated speed of sound as an approximation of the propagation speed of the finite pressure wave, the 400 mm round trip from the valve to the step and back takes about 600 microseconds, or 65 degrees of crankshaft travel.

Assume that, in an 18,000 RPM engine, the establishment of enough exhaust valve opening to allow meaningful flow would occur in the neighborhood of 100° after TDC. Therefore, it is clear that this first reflection is timed to arrive back at the valves even before the piston reaches BDC. For what purpose? Recalling that during blowdown, there is sufficient pressure ratio in the cylinder to establish choked (sonic) flow through the exhaust valve orifice, then it would certainly be advantageous to maintain that gas velocity for as long as possible.

A noted engineer in the world of Formula-One confirmed that this is exactly the reason for the one-or-more large-magnitude steps in the primary: to place a negative pressure at the back of the exhaust valve timed so as to extend the duration of the critical pressure ratio.


The required Cup engine configuration (90° V8 with a two-plane crankshaft) provides an interesting challenge for exhaust system designers. Because of the firing order of this engine configuration, the exhaust pulses on each bank of the engine are unevenly-spaced. In the words of the technical director of one prominent team: „The exhaust system design in Cup is an interesting tradeoff between minimizing flow losses while at the same time trying to optimize whatever tuning you can do with a non-equally-spaced system, which isn’t a lot.“

With the GM cylinder-numbering system (1-3-5-7 on the left) and firing order {18436572; the 4-7 swap is not allowed in Cup} ), the exhaust pulse spacing on the left side (expressed in terms of degrees of crankshaft rotation) is 270°-180°-90°-180° while the spacing on the right side is 90°-180°-270°-180°. This uneven pulse spacing gravely impedes the achievement of a well-tuned exhaust system such as can be achieved with evenly-spaced pulses and a 4-into-1 collector.

That tuning difficulty led (more than a decade ago) to the re-introduction of the 4-into-2-into-1 (so-called „Tri-Y“) configuration, which has been around since at least the 1960’s. In the „Tri-Y“, cylinders on each bank are paired so as to provide the maximum separation between pulses. Using the above numbering scheme, the primaries of cylinders 1 & 5 and 3 & 7 would be merged into slightly larger secondary pipes, which after the appropriate length, would be merged into the larger collector. On the right side, adjacent primaries are paired (2 & 4, 6 & 8). That provides a 450°-270° separation between pulses in each secondary. An example of this configuration is shown inFigure Six.

Example of a 4-2-1 Header System

Figure 6
Example of a 4-2-1 Header System

The tuning of this type of system is not terribly intuitive. Several well-placed experts in Cup told me that their teams have consumed large amounts of modeling time using very sophisticated (and expensive) simulation software to arrive „in the ball park“, and then fine tune the designs on the dyno. And, as would be expected, there are different header designs for long tracks, short tracks, and restrictor-plate tracks.

One expert mentioned that, while it is relatively straightforward to accurately model the behavior of the primaries, it is very difficult to accurately model the secondaries and collectors, because the theoretical reflections are meaningfully altered by specifics of geometry (bend radii, intersection angles, nozzle and diffuser angles, etc.) which cause destructive interference and pulse attenuation. That being said, several experts agreed that the rules-of-thumb still apply: better low end needs smaller and longer tubes; better high end needs bigger and shorter tubes.

There are additional challenges in Cup header design and tuning. The chassis teams often impose a major set of constraints on primary length and bend location so as to not interfere with critical items such as upper control arm pivot locations. The prevailing view is that, in terms of lap times, making the car turn better is a reasonable trade-off against a small amount of power increase. The straightforward header shown in Figure Six simply to illustrate the concept, is a dyno header, built almost without regard for any packaging constraint. Consider how difficult it might be to implement that concept within the very tight engine compartment of a Cup car, constrained by intruding frame tubes, suspension pickup points, a 230-mm long external oil pump, and the like.

Given the existing packaging constraints, it is indeed fortunate that the primary lengths in the 4-2-1 system are not nearly so critical as are the lengths of the secondaries. Several experts told me that the engines are very sensitive to changes in the length of secondary sections, and that most of the development effort is focused on secondary merge, length, diameter and step issues.

Moving rearward, the NASCAR Cup rulebook provides some interesting insight into additional exhaust system challenges. The rules include the requirements that the exhaust system for each bank of the V8 engine must be completely separate and may not connect in any location except for a single „X“ or „H“ pipe in a tightly-constrained region of the tailpipes, and must end with two tailpipes which exit under the frame rails within a tightly-constrained area on the right side of the car. Further, the pipes from the collector to the exit must be magnetic steel, no larger than 101.6 mm (4.0 inches) ID, and may have a circumference no greater than 336.5 mm (13.25″).

The circumference restriction provides a subtle challenge. In order to fit beneath the COT frame and still provide ground clearance, the large diameter tailpipes are reshaped into a cross-sectional form having two long parallel walls (no closer to each other than 51 mm ) and a full radius at each end, such as illustrated in Figure Seven.

NASCAR Under-Frame Exhaust Pipe Exit

Figure 7
NASCAR Under-Frame Exhaust Pipe Exit

Because of the fact that a circular section provides the most cross-sectional area for a given circumference, the necessary ovalling of the pipe exit puts an orifice at the end of the tailpipe. If the exit section is the minimum height of 51 mm, the limiting circumference (assuming 1.6 mm-wall tube) yields a cross-sectional area which is only 77% of the 101.6 mm round tailpipe. That reduced area can be a flow restriction at high RPM.

Top Fuel and Funny Car

At the top levels of drag racing, in particular Top Fuel and Funny-Car, the exhaust systems might seem very simple. The header systems, known as „zoomies,“ consist of a single pipe on each cylinder, dumping straight into the atmosphere, with each tube bent so that it faces upward, rearward, and often outward. The outward angle of bend in Funny Cars is typically larger than would be seen in an unbodied Top Fuel car, in order to eliminate bodywork damage from both temperatures and exhaust concussion forces.

In addition to the noise, a notable feature of these exhaust systems is the large volume of open, whitish flame standing just off the ends of these pipes, as shown in Figure Eight. That flame-front is the byproduct of two intersecting parameters.

Secondary Combustion

Figure 8
Secondary Combustion

First, these highly-supercharged, nitromethane-nourished engines have fuel flow rates stated to be in the 80 to 90 gallon-per-minute range. With that amount of fuel being delivered, it is clear that there will be a certain amount of fuel puddling behind the intake valve. When the intake opens, some portion of that collected fuel will be either in liquid form or in a mixture which is too rich to burn (insufficient oxygen molecules). Further, these engines apparently use a large amount of overlap in order to assist in cooling. The combination of the excess fuel and the long overlap assures that a non-trivial quantity of raw fuel and fuel mixture is short-circuited directly down the exhaust pipe, and heated during its journey. When it exits the primary, it finds an abundance of oxygen and initiates an energetic secondary combustion. The combination of the large momentum-change of the mass flow through the engine, plus this secondary combustion has been calculated by at least one aerospace engineer to generate normal reaction forces in excess of 2500 pounds (1130 kg).

Given that the pipes are angled in both the lateral and longitudinal planes, that exhaust reaction force can dramatically affect the vehicle stability. The vertical component obviously provides downforce to the chassis. The rearward component will add propulsive thrust. If everything is in balance, the sideward components generated by the left and right sets of pipes should counterbalance and net to near zero. However, I was told that the loss of one cylinder on a Funny-Car can cause the driver to have real difficulty controlling the car. That is because the loss of a single cylinder unbalances the sideward-thrust and adds a yaw-moment from the now-asymmetric rearward thrust. That same (highly-credible) source told me that the loss of two cylinders on the same bank will amost certainly render the car uncontrollable.

As might be expected, the length of the primaries plays a critical role in the engine tune. I was told by a lead engineer on a prominent Funny-Car team that there was a considerable amount of development effort required just to get the gasses out from underneath the Funny-Car bodywork.

That source also said that when they tried collector-systems, the result was that the engines ran „horribly“. The theory is that the huge amount of exhaust gas flow into a relatively-confined space raised the collector pressure enough to create a destructive blockage in the collector pipe.

As far as the pipes themselves are concerned, it is well known that „too sharp a bend“ in the primary or „too much length“ dramatically reduces engine performance. Apparently, in supercharged nitromethane engines, any tuning on the exhaust side (cam, ports, headers) requires a substantial alteration in the fuel delivery curves. After experimenting with various exhaust system changes, then working to get the fuel system back into line with the engine changes, the net change in performance was typically considered to be not worth the time and effort. After having determined a working combination, experience has shown that development efforts in areas other than the exhaust system will be more productive.

I was told that currently, there is not a large amount of development effort on the Funny Car exhaust system, as the result of several practical and economic factors. It is hard to imagine the level of difficulty involved in doing engine development on a system which is not well suited to a dyno cell, and therefore must be tested on the track in 5-second test sessions. Without taking into account salaries, logistics, transportation, food, lodging, and other „overhead“ expenses, the out-of-pocket cost to make „one more test run“ is uncomfortably close to ten thousand dollars.


Neil Spalding, Race Engine Technology’s in-house expert on motorcycles, provided me with a gallery of detailed photos showing the varied strategies employed in Moto-GP (the F-1 of motorcycle racing) to shape the engine power curves with exhaust tuning finesse, along with a wealth of information on these machines, including the fact that the use of Inconel tubing is fairly common.

In several RET articles, Neil has discussed the difficulty in getting the available power to the ground in Moto-GP, and the efforts which the manufacturers have taken to improve the available traction, including implementation of uneven firing orders so as to affect the tire contact patch in a beneficial way. The uneven spacing of exhaust pulses requires some out-of-the-box thinking to gain benefit from exhaust tuning. In order to linearize the engine power curve (flatten the torque curve) there has been widespread usage of the 4-2-1 design described above in the Cup section.

These systems use various techniques specific to the particular engine, including diverging tapers in the primary tubes just past the flange, steps in the primary tubes, converging-diverging collectors, straight collectors, diverging tapered collectors, and more.

Figure Nine shows the torturous 4-2-1 system developed for the 2005 Yamaha 990 cc irregular-fire inline 4. The picture shows the diverging taper in the primary just past the flanges. Neil told me that the current system for the 800-cc engine has substantially shorter primaries and secondaries due too the fact that the 800 cc engines turn up to 18,000 RPM, where the 990’s were in the 16,000 RPM range.

2005 Yamaha 990

Figure 9
2005 Yamaha 990

Figure Ten shows the individual stacks used on an experimental Kawasaki 990 cc engine, which reportedly had a flat-plane crankshaft but which fired pairs of cylinders together. Note the very long tapered expansion pipes and reduced exit diameters, which will help reduce the extreme power-curve peakiness that occurs when a primary opens directly into the atmosphere (which constitutes an apparent infinite expansion area ratio). Note also how the lower tube has a longer centerline length and a longer tapered end. That too will help to spread the potentially very peaky tune of these pipes over a wider RPM band.

2005 Experimental Kawasaki 990

Figure 10
2005 Experimental Kawasaki 990

Turbocharged Applications

According to the turbocharger engineers, the most important aspect of designing a good header system for a turbocharged application is to maximize the recovery of exhaust pulse energy. This energy recovery has at least two components.

The first is to provide evenly-spaced exhaust pulses to the turbine. To accomplish that, it is helpful first to be working with an engine (or bank of an engine) which has evenly-spaced firing intervals. In an application in which the cylinders feeding a given turbine or turbine section have even spacing, the lengths of the primary tubes should be as close to equal length as possible.

The second component is to maximize the recovery of pulse velocity energy. For that purpose, turbine housings are available in split housing, or „twin-scroll“, configurations, in which there is a divider wall in the center of the turbine nozzle housing to separate the incoming flow into two separate streams. That allows the nearly ideal pulse separation of 240 crankshaft degrees to be achieved on an inline-6 engine by grouping the front 3 cylinders into one side of the housing and the rear three cylinders into the other side. The same effect can be achieved on a V6 engine by grouping each bank separately.

Although the split housing arrangement adds wetted area (hence boundary layer drag) to the gas flow, the advantages more than offset that drag increase. In instances where pulse energy recovery has been optimized, it is often possible, based on calculations using pressure and temperature losses across the turbine, to observe very high turbine efficiencies, which some experts say are in excess of 100%.

Pulses which are evenly-spaced but too close together will reduce the effectiveness of this pulse energy recovery. Apparently, that phenomenon is seen on even-fire inline 4-cylinder engines as well as on individual banks of flat-crank V8 engines, where the pulse separation is 180°. I was told that the ideal pulse separation was in the neighborhood of 240 crankshaft degrees, and that, on an even-fire (single-plane crankshaft) inline-4 (as opposed to the two-plane crankshafts used in some Moto-GP motorcycle engines) it is better to separate the end cylinders into one side and the center two into the other side of the turbine than to run all four together into an undivided scroll housing. The same reasoning applies to each bank of a V8 with a single-plane crankshaft.

With regard to the uneven pulse spacing of each bank of a two-plane crank V8, there is agreement that it is very difficult to organize the pulse spacing in a useful way. It has been demonstrated that where a small turbo is used on each bank, the use of a short-tube 4 into 2 system (same idea as the 4-2-1 discussed above) feeding a twin-scroll turbine could take some advantage of the resulting 450 – 270 separation in terms of pulse energy recovery. If a single, large turbo can be located in such a way that the tubing lengths from each bank can be fairly equal, then splitting the primaries to achieve 180° separation would be an advantage.

Whenever practical, reducing the heat (energy) losses before the exhaust gasses reach the turbine allows the turbine to be more effective. This has been done with double wall tubing, reflective coatings, and wraps. However, insulating the pipes to reduce heat loss will, of course, raise the operating temperature of the pipes themselves, which can require extreme-duty materials where more affordable materials would suffice in the un-insulated form.

Another important exhaust system consideration, in order to provide the most effective operation of the wastegate in controlling boost, is to position the wastegate inlet port so that it is subject to exhaust stream total pressure, rather than off to the side where it sees only static pressure.“






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calculating header design

Discussion in ‚Exhaust and Mufflers‚ started by grumpyvette, Sep 20, 2008.

  1. grumpyvette

    grumpyvetteAdministratorStaff Member

    if your serious about maintaining good peak hp numbers use the calculators in the linked threads to actually calculate the ideal matched header dimensions, this is not a guessing game its strait physics and easily calculated to maximize exhaust scavenging and max volumetric efficiency, resulting in max power , at any given rpm and displacement, compression , cam timing etc this stuff can be calculated, but when dealing with an average and without knowing all the specifics you go by averages,and assume that peak hp takes a higher value than off idle torque
    ok as a rule of thumb,when calculating and engines expected exhaust flow rates you can roughly assume 2.257 cfm per horsepower produced

    if you have a 500hp engine it will in theory produce 500 x 2.257=1128.5 cfm of exhaust flow

    or 141 cfm exhaust flow per cylinder, but your exhaust valve is only open about 220-290 degrees and the time between power strokes gets shorter as the rpms increase.
    Of the 720 degrees in a complete engine cycle, theres lets say 250 degrees , or 250/720=1/3rd of the time, but it doesn,t all exit instantly so you can figure on about 1/2 the time the header primarys under significant pressure and pressure and pulse frequency varies with the rpms so the exhaust dump rates not at a constant flow or pressure, the header must handle 141 x 2=282cfm /115 =2.45 sq inches for exhaust flow per cubic inch , you look at the chart and find thats about a 1.7/8 diam. subtracting the pipe wall thickness





    watch video


    keep firmly in mind that the header exhaust scavenging ,efficiency and intake runners ability to effectively fill the engines cylinders is very much dependent on carefully matched, cam timing, headers designed to match the engines displacement, intended rpm band, ,maximum compression ratio, for the fuel octane used,and a low restriction exhaust behind the header collectors, when all the factors are well matched correctly theres an easy 20% or higher power gain over most stock engines in this area alone.
    correctly matched headers, compression, cam timing, displacement etc, can easily increase the cylinder fill efficiency, and intake runner air flow velocity by over 30%
    adding an anti-reversion baffle to a 18″-24″ longer collector on open headers with the internal pipe about 1/2 the collector length tends to help scavenging on some engines, no header will function to full efficiency with any significant back pressure so take the effort to accurately measure any restriction to exhaust flow AT your upper rpm range of your engines power band and be darn sure its not choking your power curve.









    you will occasionally see dyno test results similar to this, that show about a 20 hp spread in the power curve, that seem to indicate that the difference in header configuration is not overly critical,but keep in mind these results are almost always done on 350-383 displacement engines with fairly mild compression , below 9.5:1 and fairly low duration cams ( example ,below 230 deg @.050 lift )and stock or mildly restricted exhaust systems, an engine with a fairly tight LSA and high compression and a low restriction or open exhaust can take full advantage of the headers scavenging the cylinders in a tuned rpm range significantly boosting the power produced, the tuned exhaust header has a greater effect on a higher compression ratio engine of larger displacement with a tighter LSA cam
    you might want to read these also



    The primary header pipe diameter is determined using basic engine mechanical specifications, such as: Bore Stroke Compression Ratio Valve diameter Cam specifications (lift and duration) Target rpm range







    http://www.autolounge.net/calculators/e … izing.html





    http://www.superchevy.com/technical/eng … index.html


    http://www.superchevy.com/technical/eng … index.html




    http://www.carcraft.com/techarticles/03 … index.html

    http://autolounge.net/calculators/exhau … izing.html

    most exhaust pipe is 16 gauge, or .065 wall thickness


    How do you judge the i.d of a pipe?

    3″ o.d = 2.87″ diam. inside
    2.75 o.d = 2.62″diam. inside
    2.5 o.d = 2.37″ diam. inside
    2.25 o.d = 2.12″ diam. inside

    theres THREE distinct areas of pressure/velocity

    gases in the cylinders and header primary tubes as the exhaust exits the cylinders

    gases in the header primary tubes and header primary tubes as the exhaust exits the header COLLECTORS

    (3) the exhaust system past the collectors

    (1) and (2) must be calculated to match the intended displacement ,cam timing and intended rpm range

    (3) (traditional back pressure ) should be MINIMIZED


    keep in mind installing an (X) almost increases the effective cross sectional area, of the dual exhaust ,or collector cross sectional area to double what it had been behind a single header collector, by doubling the area that the exhaust flow sees, dropping the restriction to flow almost in half

    looking thru an (X) pipe








    „The overall length of the primary header pipe is governed almost exclusively by the target engine’s rpm range, and displacement which is dependent upon wave tuning. Typically, a lower engine rpm range likes a longer primary pipe, while a high rpm engine prefers a shorter primary.“

    SECONDARY PIPE DIAMETER (collector length in a 4 into one design)
    While typical off-the-shelf street 4-into-1 headers do not have true secondary pipes, Burns‘ research has proven repeatedly that his Tri-Y designs make more overall power over a broader rpm range. While traditional lines of thought have street enthusiasts knowing Tri-Y pipes make more bottom-end torque, further research by Burns into the design have resulted in headers making more power all across the rpm range. With more components as part of the Tri-Y design, more tuning possibilities exist, and therefore more potential lives within.

    „The secondary pipe diameter is determined by considering both pressure waves and reflective waves throughout the system. Since the pipes are paired according to the firing order, these waves can work together or against each other. Naturally, our designs work with the waves to increase the efficiency of the header, using the wave pulses to help pull gases from the engine.

    „There are two basic kinds of waves we’re dealing with. First, there are pressure waves. The pressure wave travels the length of the primary pipe in a 4-into-1 header, then is reflected from the collector where the area changes from the small-diameter primary into the larger-area collector. A reflection of negative pressure goes back up the primary pipe.

    „In a Tri-Y design, the pressure of additional area changes (where the primary pipes become secondary pipes) produces additional reflections, so the Tri-Y must be designed in a different manner with respect to wave control. Given this, the area of the Tri-Y header between the first and second collectors becomes critical, and tuneable. The entire header is affected by this crucial length of pipe, and can be fine-tuned accordingly through proper sizing for optimal broad-range performance.

    „The 4-into-1 pipe is also affected by altering pipe lengths, of course. But, without these secondary pipes it is impossible to tune with the same level of precision as with the Tri-Y headers. It’s for this reason we prefer the Tri-Y design in most applications. The tuneability is so much more accurate, we’re able to find more power over a broader rpm range. This is especially critical in engines expected to work well over a wide rpm range, like street machines.“

    Another huge reason for the move to Tri-Y headers is weight savings. Burn’s claims most of their Tri-Y headers weigh in at about _ the weight of comparable 4-into-1 pipes for the same application, due to the smaller pipe diameters used throughout similar applications. Also, with less internal volume than comparable 4-into-1 headers, the Tri-Y equipped engine is typically more responsive. Tri-Y designs require physically smaller collectors as well, contributing further to space and fit concerns, and adding further to crisp engine responsiveness.

    „There is much power to be found in researching collector design and size. The optimal collector is determined by several variables, and it’s engineering interacts with the entire exhaust system. The internal volume, the outlet size diameter, and the angles at which the pipes come together within the collector are all factors that must be maximized for the header to perform to its full potential.“

    1 – Primary Pipe Entry Size
    „Our computer model design program determines many of these hard dimensions based on data gathered over many years, including the length and diameter of the primary or secondary pipe entering the collector.“

    „The pipe entry angle is typically between 10-20 degrees, with most pipes being right at 15 degrees. The cone (or goilet) formed between the pipes as they transition from primary to collector is formed as a consequence of these angles, nothing more. The mass of gases moving through the pipe does not want to change direction, so keeping these „pyramid“ cones true to the pipe entry angle helps smooth the transition from the relatively small volume of the feed pipe to the larger volume of the collector.“

    „The collector outlet diameter is the most critical dimension in the header. It’s what makes the merged collector work the way it does. Each collector we sell is custom-sized to each customer’s engine, and there’s no real ‚formula‘ to get a broad-based general determination for street machines. As a rule, the overwhelming majority of aftermarket headers designed for the street market have way too big of a collector outlet diameter. Most street guys are losing power because of badly designed, manufactured, or engineered street headers. There is much room for improvement here.“

    „Overall collector length is not critical. Once the other variables in the header design have been determined, the collector ends up being as long as it needs to be. We’ve found no benefit in lengthening or minimizing this dimension. It’s more important to properly engineer what’s going on inside the merged collector, and let the length determine itself once all the other important factors are optimized.“

    One of the growing areas of research at Burns is the critical area just aft of the all-important collector outlet. Burns‘ dyno research led him to begin experimenting with interchangeable venturis, which slip into receivers just aft of the collector. While these prototype dyno parts were initially crafted to assist Jack in determining the critical overall collector diameter size, he soon realized they could be a marketable product. His initial „DynoSYS“ product for dyno research evolved into a line of interchangeable sleeves called the Burns Tuneable Exhaust Collector, or BTEC for short.

    The BTEC system has shown capability to alter the entire power curve of the engine. By changing only the insert, racers can change the entire tune on their engines to fine-tune for track conditions, weather, or driver preferences. Mostly used by drag racers, many in Pro Stock, the BTEC system offers enthusiasts a glimpse into the future of header design. While the drag racers have already embraced the benefits of BTEC, a number of road racers are beginning to experiment with the system as well.

    One area street machine enthusiasts are aware of is the evolution of the X-pipe. Early on, connecting the left and right halves of a true-dual exhaust system with an H-pipe resulted in measurable benefits. This theory evolved into the X-pipe, which allowed both left and right portions of the exhaust system to share some common flow area and resulted in even greater gains in power with a notable reduction in exhaust noise.


    placing two x pipes in sequence seems to work well at both mellowing the exhaust note, and increasing the exhaust scavenging of the cylinders as it blends and smooths out the exhaust flow by allowing the individual cylinder pulse strength to dissipate rapidly, the first (x) reduces flow restriction, the second allows the exhaust pulse to run into itself further disrupting the individual pulse strength


    http://www.engr.colostate.edu/~allan/fl … /pipe.html

    http://www.engr.colostate.edu/~allan/fl … age7f.html






    at lower engine rpms less ignition advance is needed because theres more time available, between ignition and cylinder pressure building , over the piston ,as the flame crosses the cylinder, so most of the pressure occurs after the cranks rod journal passes TDC, at lower rpms this burn & pressure build can take 50 thousands of a second, as rpms increase the time available is much shorter requiring a longer lead time or a greater „ADVANCE“ but as rpms further increase ,turbulence caused by rapid compression increasingly speeds burn times



    http://www.carcraft.com/techarticles/he … index.html








    http://racingarticles.com/blog/2008/02/ … aders.html

    Last edited by a moderator: Feb 6, 2016

    grumpyvette, Sep 20, 2008


  2. grumpyvette

    grumpyvetteAdministratorStaff Member

    theres several good calculators and extensive info you can use in these links










    http://www.otter-ag.ee/files/Header design.pdf









    http://www.pontiacstreetperformance.com … aust3.html

    http://www.popularhotrodding.com/tech/1 … index.html

    http://www.speedwaybids.com/forum/viewt … 6e4152fd32


    don,t think your MAX rpm is what you use to compute, use the AVERAGE rpm durring a run








    youll need this








    Last edited by a moderator: Sep 12, 2015

    grumpyvette, Oct 27, 2008


  3. grumpyvette

    grumpyvetteAdministratorStaff Member




    http://www.carcraft.com/techarticles/he … usion.html


    http://www.spectrum5racing.com/Technica … erRevA.pdf

    your engines cam timing , compression ratio, displacement, rear gearing and average rpm , intake manifolds port design and cross sectional area, all effect the correct choice.
    but if your into building a true custom set of headers there ARE advantages to getting the size and length of the components correctly matched to the engine.

    Ive found that if your header is designed too work peak efficiency at about 750-1000rpm under your red line youll generally be close to ideal on a street / strip combo.

    lets say that 468 BBC runs in the 3500rpm-6200rpm band when racing, your cam, timing and compression ratio, matches and the intakes designed to that rpm band.
    play with the calculators and for a street combo your probably going to want a header designed for both mid rpm tq and low high rpm restriction, a compromise of about a 2″-2.125″ primairy about 36″ long and a 3.5″ collector about 18″ long , seems about correct to maximize the mid rpm torque (remember the headers need to be fabricated from easily available components)

    ON a 383 sbc, youll probably want a 1.75″ about 38″ long and a 3″ collector about 19″ long to maximize the mid rpm torque

    grumpyvette, Dec 21, 2008


  4. grumpyvette

    grumpyvetteAdministratorStaff Member


    for you guys that refuse to read thru more than basic info links

    BTW youll need your cam cards timing and/ maybe this chart, IF the advertised durration is all you have..


    EXAMPLE LETS ASSUME YOUR USEING A CROWER 01297 cam in a 496 displacement BBC and your trying for a max of 6000rpm
    the cam timings given at .050 lift and we want the figures at about .004 so lets just assume we add 15 degrees to the timing and well be close enought to ballpark the exhaust size, needed.

    http://www.crower.com/misc/cam_spec/cam … &x=28&y=14


    Your Primary Tube Length is 37.94
    Your Primary Tube Diameter is 1.99 inches
    Your Collector Length is 18.97 inches
    Your Collector Diameter is 3.79 inches .

    THAT WILL RESULT IN A HEADER THATS A DECENT COMPROMISE that provides a good power curve.


    most manufacturers are far more concerned with low manufacturing cost and having a SINGLE or at least a few limited, and nearly universal designs,being produced, that cover as many different applications as possible to limit the required inventory to as few part numbers as they can to cover as many applications as they can.
    thats one reason the primary tube length and collector length on many commercial headers are shorter than they would be if the ideal header were calculated for every application. add to that theres several dozen different cylinder head designs and spark plug locations, different oil pans, different starter designs, different k-frames or cross members and as bad as they do fit its frankly a damn miracle, they fit as well as they do at times, and its also why any serious hot rodder needs a decent MIG, or TIG welder and the skill to use it

    grumpyvette, Jan 12, 2009


  5. grumpyvette

    grumpyvetteAdministratorStaff Member

    grumpyvette, Feb 1, 2009


  6. grumpyvette

    grumpyvetteAdministratorStaff Member

    grumpyvette, Feb 1, 2009


  7. grumpyvette

    grumpyvetteAdministratorStaff Member

    If you can,t find a decent quality header there are ways you can build a custom set-up, just a suggestion………build a custom set exactly to the correct length for your application, with the correct collector.

    BTW, its not really all that rare to find that the headers you have present a spark plug wire clearance issue,




    that makes installing the spark plug boots in such a way that they don,t contact the hot metal surfaces almost impossible , this can sometimes be helped a great deal by the addition of a header flange, or SPACER PLATE, that MATCHES YOUR PARTICULAR ENGINES CYLINDER HEAD EXHAUST PORT AND HEADERS,



    thats significantly thicker being welded to the existing header flange , to space it out further away from the cylinder heads, obviously you don,t want to do this without testing all the clearances so installing the extra header flange with an exhaust gasket on both the cylinder head surface and between the header flange spacer and the existing headers as a test is strongly suggested as a test.




    heres where you get spacer header flange plates








    Accel Extreme 9000 Ceramic Wire Sets


    ceramic plug boots are a good idea when clearances are tight


    ACCEL now offers the cure for burnt spark plug wire boots with Extreme 9000 Ceramic Wire kits. The 8mm Ferro-Spiral core wire now has ceramic boots on the spark plug end of the wires that will withstand up to 2,000° F. If you are running headers with close tolerances, an engine bay with little room, or an RV with boots that melt because of heat, these wire kits are the answer. and yeah! youll still need the heat resistant plug boots to prevent the wires from melting past the ceramic plug boot
    you start by clamping the collectors where you want them after measuring to allow room for the correct primary tube length, and bolting the header flanges to the heads with the engine in the car, read thru the inks for ideas

    theres readily available calculator info
    theres pre-cut header flanges



    theres pre-welded collectors, its really not all that difficult


    http://www.superchevy.com/technical/eng … index.html


    cutting your commercial headers short collector off and welding in a 4,2,1 extended collector is usually a good way to extend and increase the engines torque curve, but think it thru, theres the expense of the extended collectors, clearance issues and your welding and cutting and fabrication shills, I think the answer as to what course you take here depends on both your fabrication/welding measuring and cutting skills, experience and the quality of the headers your starting with, and your desire to try too maximize the engines power potential, I,m my experience theres no question, that if its done correctly youll make a bit more power, but it will also take some effort and might or might not be worth your time for the power youll gain.
    now that might be anywhere from 15 hp to 50 more hp, dependent on the combo those headers are matched too.
    and I don,t see the process of carefully fitting,the headers on the car, measuring and cutting the headers and cutting , trial fitting and welding the extended collectors several times (from experience) taking less than 4-6 hours, at 15hp that may be a waste of time in most guys way of looking at the time spent vs value gained, at 50hp i think most guys will see it as a good move, but you won,t know what youll get until your done testing, and in some cases clearances under the car make the modification and resulting under the car road clearance a total P.I.T.A., making you wish you were never involved so measure very carefully and don,t ignore the road to collector clearance issues!












    http://pitstopusa.com/i-5056669-schoenf … chevy.html


    watch this video,demonstrate, high speed air flow in adjacent primary tube sucks air flow thru next primary

    Last edited by a moderator: Nov 4, 2016

    grumpyvette, Mar 23, 2010


  8. grumpyvette

    grumpyvetteAdministratorStaff Member

    gary posted this info, I think theres a few typos, in the math , but it looks interesting

    „Ever wonder which is best 2.5 or 3 inch pipes? What about balance tubes? Is there a „right“ place to put the muffler? Are tailpipes the hot ticket? And how do I choose or build the best header for my engine?

    Get your calculator out, put your feet up, and start designing your own header and exhaust system, for your specific combo.With these exciting, ok, tedious, formulas straight from various engineering texts.

    I use these for my stuff and have seen a 6 hp gain across the board and 10 at peak on a chassis dyno when replacing the best of the pre-bent X pipe stuff with a system designed like this….

    1. First calculate individual cylinder volume.

    (((Bore x 2.54) x (Bore x 2.54)) x (Stroke x 2.54) x 6.2832) / 8 = individual cylinder volume

    2. Select desired peak horsepower rpm (i.e. 6000 RPM, from your dyno sheet if possible)

    3. Identify exhaust valve opening BBDC (Before bottom dead center) in degrees (i.e. 67 degrees BBDC, from your cam spec sheet)

    4. Calculate ED. Add 180 to BBDC spec (i.e. 180+67= 247)

    5. Identify collector diameter inlet (almost always 4 inches)

    6. Identify collector diameter outlet (almost always 3 inches)

    7. Collector cotangent angle of taper will be 7 to 8 degrees for most headers

    8. Calculate header primary tube length.

    (ED x 850) / (Peak power RPM – 3) = primary tube length in inches.

    9. Calculate primary tube inside diameter (ID)

    Square root of ((Individual cylinder volume / ((Primary length + 3) x 25))) x 2.1 = primary ID in inches

    10. Calculate collector length.

    ((Inlet diameter – outlet diameter) / 2)) x cotangent angle of taper = collector length in inches.

    11. Calculate intermediate (collector to muffler)pipe ID

    Square root of ((Individual cylinder volume x 2)/((Primary tube length + 3) x 25)) x 2 = pipe ID from end of collector taper to front of muffler

    12. Calculate intermediate pipe length

    (Primary tube length+3)-collector length = intermediate pipe length in inches

    13. Calculate balance tube ID

    1.5 x Primary tube ID = balance tube ID in inches

    14. Tailpipe. Same ID as Intermediate. Minimum 8 (12 is better) inches long to scavenge the muffler. It can be any length as long as it’s not shorter than 8 inches.

    The header you just designed will ensure that the sonic reversion pulse travels back up the pipe to the exhaust valve precisely at the right time at peak power RPM. The exhaust system will support the pulse timing and scavenge the header and muffler. See, it’s simple…“

    grumpyvette, Jul 9, 2010


  9. grumpyvette

    grumpyvetteAdministratorStaff Member

    the question often comes up about use of mandrel bends vs crimp bent exhaust pipes, in designing an exhaust system, well it should be noted that its the cross sectional area much more than the shape of the pipe thats the more important factor, while its true that mandrel bends do maintain a more consistent cross sectional area, simply selecting a slightly larger diameter non-mandrel bent exhaust pipe size with its larger cross section can frequently be the less expensive route. as long as you’ve got an (X) pipe in the system , mounted as close as clearances under the car allow, to the header collectors and the tail pipes are nominally the same diameter, as the formulas suggest are required, IE lets say 2.5″ or 3″ the type of bend at that point, (past the (X) PIPE, will be all but meaningless due to the fact that by that point the exhaust pulse strength and velocity has been significantly reduced thru cooling distance, the effect of the (X) pipe splitting the pulse,and the lack of significant restriction.
    every test Ive ever seen shows that an (x) pipe mounted near the header collectors and mandrel bends on collectors do help flow, but youve effectively almost doubled the cross sectional area after the (x)and because the engine fires every 90 degrees the pulse of exhaust past the (x) is significantly reduced in exhaust pressure, your exhaust will normally require an exhaust pipe that will handle the flow based on the engines air flow rate and horse power
    you can use the info posted



    knowing a few constants in engine pressure and flow helps

    an engine usually requires approximately 2.257 cubic feet per minute per horsepower to maximize intake flow and exhaust flow at about 115 cfm per square inch, that holds basically constant wither your spinning a 302 to 7800-8000rpm with 4000fpm in piston speed or a 406 to 6000- 6200rpmrpm

    so assuming your building a 500 hp engine / 2 (divided by 2 as there’s normally two header collectors on a v8) we have 250hp per header collector, (open header collectors) multiply that by 2.257 cfm and you see you need 565 cfm and divide that by 115/square inches and we see we need a 4.9 square inch minimum exhaust collector pipe, per side (open header collectors).

    as a cross check 500hp /8=1129/8=142 hp per header primary , 2.257 x 142/115=2.76 sq inches 0r a header primary a bit larger than 1 3/4 and smaller than 2″ or a 1 7/8 to maximize peak hp, per header primary, but keep in mind you’ll spend most of your time below peak rpms so a slightly smaller 1 3/4″ primary on a street strip engine that sacrifices a bit of peak hp for better mid rpm torque makes sense, and once you install longer exhaust pipes and mufflers you’ll need to steep up the exhaust pipe size cross section past the header collectors or they will tend to be restrictive at the minimum size the formula predicts
    at any given rpm and displacement, compression etc this stuff can be calculated, but when dealing with an average and without knowing all the specifics you go by averages,and assume that peak hp takes a higher value than off idle torque
    ok as a rule of thumb,when calculating and engines expected exhaust flow rates you can roughly assume 2.257 cfm per horsepower produced

    if you have a 500hp engine it will in theory produce 500 x 2.257=1128.5 cfm of exhaust flow

    or 141 cfm exhaust flow per cylinder, but your exhaust valve is only open about 220-290 degrees and the time between power strokes gets shorter as the rpms increase.
    Of the 720 degrees in a complete engine cycle, theres lets say 250 degrees , or 250/720=1/3rd of the time, but it doesn,t all exit instantly so you can figure on about 1/2 the time the header primarys under significant pressure and pressure and pulse frequency varies with the rpms so the exhaust dump rates not at a constant flow or pressure, the header must handle 141 x 2=282cfm /115 =2.45 sq inches for exhaust flow per cubic inch , you look at the chart and find thats about a 1.7/8 diam. subtracting the pipe wall thickness

    you might want to read these also







    you can also use intake port flow as a rule of thumb, ie number of cylinders x max port flow art max cam lift x .257 = max theoretical hp an engines likely to produce, before the heads and intake become a restriction

    IE if a vortec head 350 flows 229 cfm at .600 lift and your spinning the engine fast enough too maximize the port flow rates,so in an ideal world you can expect 229 cfm x.257 x 8 cylinder =470hp before the heads become a restriction if everything else is perfectly tuned


    grumpyvette, Mar 11, 2011


  10. grumpyvette

    grumpyvetteAdministratorStaff Member

    It ain’t just pipes, it’s science! Part 1

    I’d like to try to explain some basic exhaust theory and clear up some
    issues that may not be completely clear.

    Everyone knows the purpose of an exhaust system is to provide a means for
    the exhaust gases to be removed from the cylinder. You might wonder why
    have an exhaust system at all. Other than the frequent need to muffle the
    noise of rapid combustion, why not simply open the exhaust port to the
    atmosphere, thereby saving both weight and expense?

    Some time back in early internal combustion engine history, it was
    discovered that attaching a length of pipe to the exhaust port (probably
    to direct the noxious exhaust fumes away from a passenger compartment or
    out of a room where a stationary engine was housed) often had an effect on
    the performance of that engine. Depending on parameters such as pipe
    diameter and length, the performance could be adversely or positively

    I expect it was clear from the very beginning that exhaust gases have
    momentum. What may not have been known at the outset is that they also
    exhibit wave properties, specifically those of sound. Both those
    properties can be utilized to evacuate the exhaust gases more quickly and
    completely. The usual term for this removal process is “scavenging.”

    There are two types of scavenging: inertial and wave. Inertial
    scavenging works like an aspirator whereby some of the kinetic energy of a
    moving fluid stream (air, water, etc, generally in a pipe) is transferred
    to the fluid in an adjacent pipe. You may remember from high school
    chemistry lab class where you used water traveling through the top of a
    “T” fixture to draw a quite powerful vacuum in an attached vessel.

    The “T” can be likened to a merge collector as used in virtually all
    successful racing cars (although often not in dragsters). The most
    effective merge collectors minimize the volume increase at the juncture of
    the pipes. If this volume is too large, gas speed is diminished and less
    kinetic energy is transferred to the gases in an adjacent pipe. Thus, the
    scavenging is less complete. Well, so what if there is a little gas left
    in the pipes? Consider the engine cylinder as an extension of the exhaust
    pipe. A cylinder with residual exhaust gases has less room available to
    accommodate the incoming charge of gas and oxygen. Obviously, the more
    gas and air you can get into a cylinder, the more power is developed; that
    is why superchargers are so effective.

    Not only can scavenging be utilized to empty the cylinders, it also can
    help to draw in the new charge, by producing a negative pressure in the
    cylinder. This gets tricky because there has to be adequate time in which
    both the intake and exhaust valves are open, and there is the potential
    problem of the new charge passing right through the cylinder into the
    exhaust pipe! Gas is wasted and power is lost. Maybe you can design your
    cam such that it closes at just the right time to prevent this from
    occurring. Or maybe you can make the exhaust pipes just the right length
    so that the reflected sound waves (at a particular engine speed) prevent
    the incoming fuel and air from spilling out of the cylinder. More on this

    A stock S4 engine has very little valve overlap (some at small valve
    openings) and therefore there is only a short time during which scavenging
    of the cylinder can be accomplished. Even still, there is opportunity for
    significant performance gains with effective scavenging of the primary
    exhaust pipes (the first pipes that emanate from the ports) where it’s
    possible to produce a negative pressure so that when the exhaust valve
    opens, exit speed is increased. The result is increased momentum and
    possibly improved cylinder evacuation.

    On to wave scavenging. An analogy would be tuned organ pipes in which
    their length is adjusted such that a standing wave of a particular desired
    length (and frequency) is established. This means that some whole number
    of waves will fit exactly within the length of the particular pipe. When
    the point of maximum amplitude of a wave comes to the end of the pipe or a
    change in diameter, the wave is reflected back up the pipe, but as its
    mirror image. Thus a positive pressure wave is reflected as a negative
    pressure, or rarefaction, wave which, in turn, helps to draw spent gases
    from the pipe/cylinder. Wave scavenging is most effective over a narrow
    speed range that can be adjusted by changing the primary pipe length.
    Thus a torque or power peak can be designed to occur at a particular
    engine speed to suit the application whether it is racing or everyday

    What are crossover headers? There are numerous types of headers, tri-Y,
    equal length, stepped, unequal length, crossover, etc. Unequal length
    headers are by definition not tuned at a specific rpm; rather each pipe is
    tuned for a different speed. They tend to perform better than the stock
    manifold and may increase performance over a broad speed range. Because
    of their unequal length, each pipe will utilize wave scavenging at a
    different speed, thus reducing the effect at any single or narrow band of
    speeds. They often have sub-optimal merge collectors and so, do not make
    the best use of inertial scavenging. Equal length headers can be
    excellent wave scavengers, but often have inferior collectors, so inertial
    scavenging is not optimized. The tri-Y design is especially good on
    4-cylinder engines and is now being used almost exclusively on NASCAR
    engines with 8 cylinders. Stepped headers gradually increase the pipe
    diameter going away from the port. I believe at least one of the purposes
    is to inexpensively approximate a megaphone which is the most efficient
    device for returning the pressurized gases back to the surrounding
    atmosphere. “

    grumpyvette, May 23, 2011


  11. Indycars

    IndycarsAdministratorStaff Member

    I noticed when CCing the combustion chambers that I could see the reflected wave coming back when the wave hit the other
    side of the chamber. It would easier to see this in a pan of water while using an eye dropper, you do have to look
    closely or you will never notice what’s happening.

    Or you can watch this video. Take close notice at the 2:08 min:sec mark in the video.

    Indycars, May 23, 2011


  12. grumpyvette

    grumpyvetteAdministratorStaff Member

    http://www.popularhotrodding.com/tech/1 … index.html


    How To Size Headers Pipe Sizing Chart
    How To Size Headers
    This chart works well for high-performance heads in as-cast or ported form. To determine the primary inside diameter, you will need a good ballpark figure for the exhaust flow at the peak valve lift. Using this figure, go up the chart until you reach one of the colored lines (green for street, purple for street/strip, blue for race). At that point go across the chart and read off the pipe ID required.


    How To Size Headers
    Here are some dyno test results of four different primary tube lengths on an 8.5:1 compression 350 Chevy. The shortest length at 18 inches (black curves) show the engine did not like really short primaries. But the 29-inch (green), the 32-inch (red) and the 38-inch (blue) headers all performed well with only a moderate bias toward better low end from the longer pipes.


    Here we see the effect of adding a tuned length (about 12 extra inches) to the normal stub length seen on a regular collector. The addition of 12 hp and 10 lb-ft is good, but just check out the 40 lb-ft gain at 3,600 rpm.

    grumpyvette, Nov 6, 2011


  13. grumpyvette

    grumpyvetteAdministratorStaff Member

    this can and does change a great deal with each combo but its not all that rare to find that a 14″-24″ collector extension, added to the headers collector helps cylinder scavenging efficiency, Ive repeatedly found similar gains, while tuning cars and I,m always amazed at the guys that slap headers on a car and think thats all thats required to get the engine to perform to its max potential.
    theres calculator programs that will help predict the best combo but they rarely give you the exact info, it usually takes some actual track time and tuning to find the best combo in the real world.
    Ive generally found most collectors are too large in diameter to produce the best results, and if you calculate the cross sectional area of TWO primary pipes and install a collector just a bit larger in cross sectional area you get better results



    look at the chart, as an example if your header primary is 1.75″ its got 2.4 square inches so twice that area is , nearly 5 square inches, a 2.5″ diam collector is very close, but you generally select the next size up so 3″ diameter collector extension near 18″ long would be where Id suggest starting on a header collector extension, and if you can run the header collectors into an (X) pipe in many cases it helps scavenging efficiency even further








    grumpyvette, Sep 13, 2013


  14. grumpyvette

    grumpyvetteAdministratorStaff Member

    http://www.popularhotrodding.com/tech/1 … dyno_test/

    I can,t begin to tell you how many times Ive seen the „FACT“ that 1 5/8″ headers will produce more torque than the larger 1 3/4″ headers
    or similar info posted, the TRUTH is that each cam, engine displacement and combo of components will effect your results and only testing and a bit of experimentation with collectors and cam timing , ignition timing a bit differently can change the results


    THE HORSEPOWER NUMBERS and a test of big block chevy engine testing with different headers


    Last edited by a moderator: May 15, 2016

    grumpyvette, Feb 7, 2014


  15. philly

    phillysolid fixture here in the forum

    grumpy, can we infer from this diagram that making exhaust primary and intake runner lengths for different rpms we broaden our powerband? maybe lose some peak hp but possibly pick up plenty of average tq?

    Attached Files:

    philly, Feb 8, 2014


  16. grumpyvette

    grumpyvetteAdministratorStaff Member

    Ive always tried to get the engines compression, cam timing, displacement, intake runner tuning and exhaust scavenging to fall as close as I can in the rpm band, so i can select the converter stall and rear gear ratio to maximize that peak from the torque peak to and a bit past the power peak, deliberately screwing that power curve up to broaden and lower the power produced never entered my mind, now Im sure I screwed it up a bit at times and had some parts operating a bit more effectively that other parts at some point in the rpm band but it was never intentionally done, simply because done correctly theres still a fairly wide and useful 1500rpm-3000rpm power band you gear the car for