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Content > Engines & Driveline > Parts selection for a street performance 302 Windsor engine > Mustang Forum Australia - Mustang Tech

Parts selection for a street performance 302 Windsor engine

Parts selection for a street performance 302 Windsor engine

Author: Boofhead (as known on Mustangtech.com.au).

Notice: This article is Copyright (c) 2016 to the Author known as Boofhead on Mustangtech.com.au. It has been written for the tech section for the use by members of the Mustangtech website. It is not to be copied, reproduced or provided by any other means than a link to the original copy as published on Mustangtech.com.au. I (Boofhead) provide the permission for Hybrid (Owner and administrator of Mustangtech.com.au) to publish this work on Mustangtech.com.au only.

Disclaimer: This article is provided for information and education purposes only. The author has built many engines using processes and techniques illustrated hence the information provided is appropriate and correct to best knowledge of the author though no warranty or responsibility is given if a reader decided to follow any instructions or information provided. Many of the ideas have been developed over the years through hard experience though there will be people who may disagree with the content. The author acknowledges the are often more than one approach or procedure in part or fully in selecting parts for an engine build. You will likely find discussions in the forum on Mustangtech.com.au for many of the ideas presented. The author is open to constructive criticism so feel free and post your questions, concerns and successes. My goal is simply to help remove the mysteries surrounding engine building to assist those in this great hobby of maintaining our classic cars.

Acknowledgement: The author 'Boofhead' wishes to thank the Mustangtech community. Especially the following members for providing feedback and select images for this article; trav68, cage, shaunp, mikes68, gbx78, hybrid and ozbilt.


This technical document is designed to provide the reader with a resource guide covering considerations and the processes to illustrate how to select all of the main parts in preparation to rebuild a small block Ford Windsor V8 engine. The primary engine in the small block family is the 289/302 Windsor (also later version known as the 5.0 Windsor). The 351 Windsor has a number of differences though the basic processes are identical. Where a difference exists a note will be provided to inform the reader. The other ford small block is commonly known as the Cleveland engine will not be explicitly covered though many of the points are generally valid for the engine parts selection process. See Image 1 for a picture of a first time engine build that had been successfully executed by selecting the parts and building the engine while applying many of the suggestions and procedures as documented in this article.

Image 1: A performance 302 Windsor specified and built by a Mustangtech member with guidance from the author and the Mustangtech.com.au community.

The contents of this article may at times be very technical, the goal has been to provide the details, the reasons and the processes for choosing particular parts so when combined result in a great engine that will perform to expectations. This article is designed to be read from start to finish though feel free to read any section of interest.


  1. Background
  2. Initial decisions and aspects to consider
    1. Vehicle's intended usage
    2. Budget considerations
    3. Platform Structure and Strength
    4. Can the Vehicle Stop
    5. Further considerations
  3. Planning, preparation and parts selection
    1. The block
    2. Verify if you're being realistic
    3. Crank, Rod and Piston considerations
    4. Cylinder head selection
    5. Camshaft selection
    6. Intake manifold selection
    7. Exhaust manifold selection
  4. Misc Parts selection
    1. Major fasteners
    2. Major gaskets
    3. Positive crankcase valve
    4. Fuel Pump
    5. Intake filter
    6. Stall converter
    7. Differential ratio
  5. Engine combinations
    1. 289 / 302 cid engines
    2. 347 (+351) / 363 cid engines
    3. 393 / 408 cid engines
  6. Conclusion
  7. Frequently Asked Questions


The Ford small block was first introduced in 1962 and produced in various forms and sizes in various vehicles right through to 2001. The engines are still available new from Ford as crate engine replacements where every part is available new from OEM and after market sources. The engine was designed to be light using the latest thin casting techniques and be compact to fit the new line of compact cars such as the falcon and mustang. The engine design has proven to be an excellent product while improved over time resulting in its longevity. In fact the engine has outlived its intended first replacement the Cleveland V8.

The Windsor (nicknamed based on one of the factory locations used to build the engine) was produced in the 221, 260, 289, 302 and 351. The 221, 260, 289 shared the same stroke of 2.870"while the 302 used a 3" stroke and the 351 was a 3.5" stroke. The 289, 302 and 351 have a 4" bore. Naturally the parts selected process will often result in making use of after market performance parts as required. As expected, preparation and planning is very important though often over looked. This article assumes the use of the 8.2"deck 4" bore block of the 289 and 302 family. The goal is to provide complete information, processes and methods detailing each step on how to approach and complete the parts selection process for a street performance engine aimed for your mustang or compact ford.

Initial decisions and aspects to consider

This phase of the process is key to the overall success of the resulting build. A number of questions need to be considered and answered at the same time while being realistic with the desired result.

Vehicle's intended usage

Firstly; How are you intending to use the vehicle? I can understand the desire to have an engine that has lots of power. You can brag to your friends by quoting big figures while at the same time the idle sounds like a pet dragon as it spits fire while loudly rumbling to tell all near by that it is no ordinary car. I can also fully understand the desire to have a pure almost standard street engine just to go cruising around the country side. I generally believe somewhere in between is the most common desire.

So keep in mind if you build a nasty engine and only use the vehicle to lightly cruise the streets then you may not enjoy your car which is the point of ownership. Alternatively if you build a standard engine with the desire to race at the drag strip every week then once again you will not have a result that meets your real expectations. In both cases, I would call the build a failure. Be realistic with how you intend to use the car. If your're not sure what to build, aim for a mild engine using quality parts so then you at least have the option to upgrade a few items if your desire changes toward a more powerful engine. This is an important point particularly if this is your first performance V8 in a light mustang. A relatively light vehicle with a modest level of power will provide quite a thrill when you want one and can easily become a handful. Go for a ride with someone who has a build similar to your thoughts as this experience can often confirm if your expectations and goals will be satisfied.

Budget considerations

The second question you need to answer is; What is your budget for this build? The more powerful the engine the more it will cost. For example, a power to cost graph becomes very steep toward the cost side as the power increases. Most mild performance builds do not need to be a very high cost exercise though it may cost 50% more to get the next 100HP then double for the following 100HP. Consider all costs in the budget so include not only the large major items such as heads and block machine work, there are many small items as well. They all add up so I suggest a spreadsheet is maintained with the parts list, services and costs. Make sure you include fluids, gaskets, oil, sealers, and disposables such as rags, degreaser and other such items. You need to consider engine accessories as well as you may need (or find this is a good opportunity) to upgrade items such as the alternator to power the extra load such as radio, AC, cooling fans, EFI system and many other electric items. Accessories do not need to be sore on the eyes. For example, Image 2 illustrates new accessories combined with a dress up kit on a newly built engine. Also, consideration toward using a replacement starter motor as the current one might be tired after 50 years of service, or for space considerations to fit exhaust headers or helping with extra load due to the higher compression the engine may have beyond the original. Secondary items such a radiator sizing for cooling may need to be increased as a higher performance engine will generally generate more heat. Lastly, an important aspect that is often forgotten when looking at the higher ranges of street performance is the potential need for structural changes to the car. All of these items should be included in the budget. Finally, do not be surprised or simply expect to go over the budget as it often turns out to be the case.

Image 2: A high performance 302/347 Windsor built by a Mustangtech member with assistance from the author with all of the accessories.

Platform Structure and Strength

The Mustang is approximate half a century old and frankly in that time they were not always treated all that well. Particularly before their classic status was earned. As a result you will commonly find the structural integrity of the vehicle is often not as solid as it was from the factory. The fundamental cause is the bodies do not have a full chassis hence flex leading to metal fatigue and cracks. This is potentially coupled with poor repairs and maintenance in particular the early years. Given a performance engine will place extra load on the chassis it is worth the effort to consider addressing any issues if present. As it happens Ford was well aware of these issues and with the introduction of the big block engines they provided extra support in terms of torque boxes. In 67 only one was present then in 68 and later the left and right hand side torque boxes were installed. In later models a further support piece was added to triangulate the torque box to the main frame rail. In all cases the convertible models included extra structural materials and supports, such as both of the torque boxes, as it was necessary without the roof section in place. Ford found it necessary to improve the structural integrity of these cars and as such it is suggested to consider the installation of torque boxes if your model is missing any and if they do exist verify that their integrity is sound. A high powered engine can permanently twist a cars body resulting in expensive specialist equipment to pull the chassis straight again. Installation of the torque boxes will help address this potential problem.

The long term torsional flexing, that the torque boxes are designed to strengthen against, can result in chassis issues. A common area where chassis cracks can occur is around the shock tower suspension points where the metal is layered to support and strengthen the area for the upper and lower suspension arm connection points. This area is factory spot welded onto the main shock tower where time has proven it is not sufficiently joined to the shock tower. Often you will clearly see cracking around the suspension holes and spot weld points. At the same time visibly inspect the edge of main support layer to often see the metal lift away from the shock tower. It is suggested strongly that this area is inspected and issues addressed. The correction process normally involves using a hydraulic ram to force the separated layers back flat against the shock tower then a full vertical weld applied so separation can no longer occur as Image 3 illustrates. In addition to any tower support work all cracks require to be welded up, an option on 67 and later models to add a full tower wrap which extends the main suspension support to the sides of the shock tower as shown in Image 4. This can then be welded on all sides or optionally for a factory look it can be welded from inside the tower after cutting an appropriately large hole in the shock tower.

Image 3: Shock tower welded and ground back.

Image 4: Shock tower wraps installed on a 68 ready to be cleaned up and painted.

Note: While fixing the shock towers it is a possible opportunity to complete the Shelby drop modifications that relocate the position of the upper arms.

Once integrity is returned to the shock towers it is strongly suggested to help support the whole structure by installing a one piece shock tower brace. Lastly, an optional straight Monte-Carlo bar can be used to complete the triangulation of the engine bay tying the shock towers together. The suggested repairs and modifications all work together with the torque boxes and chassis rails to support the vehicle's structure in preparation for the new engine resulting in years of safe driving and enjoyment.

Can the Vehicle Stop

The extra performance provided by an upgraded engine will often require the braking capacity of the car to be addressed. All mustangs had front disk brakes as an option hence I would consider it mandatory to have at least the factory or equivalent front disk brakes installed. In the case where the system is original then install a full service kit with fresh hydraulic fluid through the entire system. A well functioning factory style front disk brake setup will have plenty of capability to suite a street oriented performance mustang. Upgrading the brake pads, such as using the EBC Green Brake pads, can provide further brake rotor bite if it is found necessary.

Alliteratively, assuming the intention is to enter into autox, track or drag racing on a semi-frequent or certainly on a serious basis, then there are many options to provide additional braking capability. The most common first step is to install larger disks, possibly front and back. The use of slotted brake rotors is worth the small extra cost while cross drilled disks look great are generally not necessary. Larger frictional areas often found with after market calipers that employ larger brake pads offer higher braking potential due to the greater frictional surface area as shown in Image 5. There are many brake kits available as such the main consideration of selecting a kit for use on a street registered car in Australia is the calipers include dust seals. The master cylinder in most cases will need to be upgraded as well as is the possibility for the requirement of a proportional valve. Please remember that any non factory brake upgrade requires an engineers report so include this in your budget plan.

Image 5: The after market offers many options to improve braking performance beyond the factory standard.

Disk braking systems are usually supported with a vacuum based booster. Performance engines generally employ a camshaft with increased camshaft overlap that can result in a lower vacuum signal at idle that can affect the power brake booster's effectiveness. Depending on parts selection and the owners preferences, this can be an issue. An engine that is targeting a very high performance bracket the vacuum level is will likely be to low resulting in the brake power booster to become ineffective. As the owner you are the best person to judge if any greater required foot pressure is uncomfortable or possibly dangerous. In which case there are three main options. Firstly a larger booster is installed. This appears to be a simple solution though space in the engine bay and potential installation complications of a non-factory sized booster can restrict this option. The second approach is to employee a secondary vacuum pump. These devices can be obtained in either electric or mechanical forms. The third option is to use a hydro boost system. This approach taps into the power steering pump pressure to create a pressure reserve that is the source of additional brake peddle force. Most street engines do not require any changes though if it is required then any of the above options can be affective methods to provide extra boost to improve brake feel so it is comfortable and safe for the driver.

Further considerations

There are a few minor aspects left to consider. All depend on the level of performance targeted. At the front of the car is the cooling system that is likely in need of being enhance with the installation of a performance oriented engine. The 67 and later mustangs had the big block option which also included a larger radiator. This approach would be able to cool a performance small block Windsor. The early mustang can install the later radiator or employ an after market option. Careful selection of the coolant used and checking for stray voltage is important especially when using alloy components. Under the center of the car the tailshaft needs to be checked and serviced so it is in good working order. It is suggested that new universal joints are installed. Check that it is straight and in balance. A standard V8 tailshaft would be sufficient for any street oriented build. Any higher end performance build is likely to require a stronger tailshaft while it would be wise to also install a tailshaft hoop for those tracking or drag racing. At the rear is the differential that is required to transfer the engines power and torque to the rear tyres. Often an area of neglect so full service, new oil, seals and bearings are needed. While restoring the differential, the center employed as well as the ratio being used could be changed.

The differential ratio and center choice should be an integral part of the engine build plan. The target diff ratio is a part of the discussion involving camshaft choice while housing and center can be constrained by budgets. An early model 8 inch mustang diff is fine for the majority of street engines while for higher end performance it is best to move to the stronger 9 inch ford rear. In most cases, the diff center is an open diff center that are generally the most likely part to break when the car is driven hard. A limited slip or locking diff center is a better option as the design implies is it able to transfer the power equally to both sides of the diff. The toughest center to use is the Detroit Locker which was an option on many of the big block mustangs. Its design is a true locking diff center though they are not always smooth or silent in operation in tight slow corning situations. Modern designs have improved significantly the lock/unlock clicking that can occur. A true locking center is the best option if drag racing is the intended use while it still allows normal street driving to and from the track. The majority of owners would find the use of a limited slip diff design more suitable due to its quiet operation. There are two main designs. Firstly, the tradition limited slip design that employs clutch plates to provide control. Unfortunately, over time, like any clutch plate they tend to wear which results in reduced performance. A superior approach is to install a Detroit Truetrak center that employs gears instead of clutch plates to provide the smooth limited locking operational behaviour that retains is performance even with extended use.

The success of a build is often more of a result from good planning with consideration of all aspects of the goals and intended use of the vehicle.

Planning, preparation and parts selection

The parts selection process has the over all aim to successfully select the best possible parts that work well together within your budget for the performance goals being targeted. The parts will fit into two main groups; the horse power and reliability categories. As the performance goals increase the greater the funds are needed for both categories though there is a point where reliability starts to dominate the costs. Keep in mind the saying, "you can complete the build and satisfy only ever two of the following categories; Cheap, Dependable, and Fast". Choose two as you cannot have the third. Fortunately, targeting street performance there are a large range of parts from the after market manufacturers that are of excellent quality at very reasonable prices. As a result, we are able to avoid many of the more exotic parts that would be necessary for a top engine build.

The block

The very first decision is to choose the base engine you are going build. The selection of the base block should involve consideration of the budget, performance and fitment to the target vehicle. For example, for an early mustang it is best to select a 289 or 302 block as the base. The engine fits well, is light and has excellent performance potential. A 351W can be chosen though the block is one inch wider and one inch taller. As a result the space available for changing spark plugs is reduced therefore it becomes a slightly more difficult process, so does exhaust header selection and fitment as well as bonnet space for intake and carburetor with height requirements that can become an issue. In addition, be prepared for heavier steering due to the added weight. Even with the difficulties detailed the 351W is a viable and common choice especially when a stronger factory block is a motivating factor. In the case of the later mustangs, that have a wider engine bay, the larger 351W engine is practical while especially in a stroker form is a preferable choice.

Image 6: The 302 Windsor block is a light compact engine.

Over the years there have been only a few changes to the basic Windsor block. The early 289 and 302 Windsors were heavier and stronger than the later roller block castings. So I personally prefer to use an early block for an early mustang if a decent block can be sourced. The early blocks were also balanced to 28oz external weight while the later roller blocks where changed to a 50oz balance weight. The lighter the ballance weight the better. The thinner casting and heavier balance introduced in the later engines reduced over weight which, in theory, saved owners fuel while for ford savings in materials resulted in a lot of money saved in production costs over the life of the Windsor family.

Assuming a decent early block can be sourced then it is the superior engine to build. One option is to use an after market block such as one that Dart provides. While expensive, these blocks have the advantage of being compact while being far stronger than any factory block with the option of going to a larger than factory bore that provides additional engine capacity. Assuming your budget is able, this is the best choice for a high output engine in an early mustang. The point at which a Dart block should be considered is when targeting over 500HP at the crank or any level of hard abuse or a stroked engine that will be subject to sustained 7000 RPM plus. The 351W can take far higher HP limits approaching 700 HP though high sustained RPM is not wise due to the vary large 3 inch main crank journals. If purchasing a 351W Dart block then select the Cleveland crank main journal size then combined with an appropriately available stroker crank.

An additional difference is the early 289 and 302 blocks did not have any head bolt threads going into the water jacket while on the later blocks the outside (short length head) bolt holes do go into the water jacket. This necessitates the use of a thread sealant to be applied to the threads on later engine outer short length head bolts. Similar to the early 289 / 302 engines, the 351W has always retained the 28oz balance. In very high RPM applications a neutral balanced crank is best though not necessary in most street engine applications where a 28oz balance is perfectly suitable and preferred over the 50oz crank.

An oddity of the 351W family in the very early blocks having been produced with a slightly shorter deck. These are the strongest factory 351W blocks though the short deck may necessitate some careful parts selection. There are a few advantages to be found in the later blocks. Firstly, the factory support for using a roller camshaft is a significant advantage. It is suggested to verify the 351W roller block's integrity, in some cases, it may have small cracks around the oil feed to the cam bearings that are located close to the spider mount points. It will normally not be a problem though I would avoid in a significantly higher performance build. This is not a known issue with the 302 roller blocks. Lastly, there was a change in the camshaft firing order in the 302W (often referred to as the High output, e.g., HO) to be the same as the 351W. The design change reduces the front main bearing load by moving a double ignition firing event to the rear bearing that is located in a structurally stronger section of the block. The later blocks employ a one piece rear main seal that is generally less troublesome. Lastly, the later blocks are easier to source and often have minimal ware while showing low signs of corrosion due to improvements in oils and coolant additives over the last 50 years.

Verify if you're being realistic

At this stage most have an idea of the engine they want to build and the HP they would like to achieve. I hear all the time talk along the lines of "I want to build a 500 HP 289 or 400 HP 347 etc" so this is a good point to check if your goals are realistic or not. In general the larger the engine the greater the performance is for less work (meaning RPM). The size of the engine sets the performance potential in combination with the target Peak HP RPM. For example, a simple method to verify realistic targets is a generialism that states a strong street performance engine can attain 1.3 times its capacity in HP. So a street 302 has the potential for 390HP, a 347 is 450HP while a 408 has 530HP at the same maximum RPM. This is a quick way to select the engine and its stroke, e.g. if you want a 500HP street engine for the bragging rights you will need at least a 393 CID engine at Peak HP at least 6000 RPM. This illustrates a 500HP 289, although possible, is not a realistic target for a street engine.

For some people, like myself, the generalism is not a formal enough process so an alternate verification method is required. A formal formula that I use to verify engine potential involves calculating the required Volumetric Efficiency (VE) of the proposed engine. A normal stock engine would be 0.7 to 0.8 (70 to 80%), a good street performance engine would have a VE of 0.9 (90%), a very strong normally aspirated engine would have a VE of 1 (100%) while exceptional engines can be higher. The VE represent the engines ability to be filled with air to its capacity. For example, if an engine has a VE of 1.1 (110%) then it fills with 10% more air than its CID. This is how a Turbo or Supercharger works by pressurising the intake to force air/fuel into the engine artificially making the engine affectively larger. In a normally aspirated engine, a well selected performance parts combination with a high average piston speed affectively improves the VE of the engine. The formula to use is;

Formula to calculate a proposed performance engine VE.
VEVolumeric Efficiency (Required)
CIDCubic Inch Displacement
HPEngine HP
BSFCBrake Specific Fuel Consumption (use 0.45)
REQUIRED VE = (9411 x HP x BSFC) / (CID x RPM)

For example verifying a goal for a 289 to have 500HP at 5000 RPM;

VE = (9411 x 500 x 0.45) / (289 x 5000)
VE = 1.47 (so 147% Required VE)

A VE requirement of 147% is very far from possible for a normally aspriated engine. Supercharging would be required at approximately 15 psi [1 bar] boost for that power level. While a normally aspriated 400 HP 347 at 5800 RPM has a required VE of 0.84 (84%) which is realistic with the right parts selection. Once you have a feel for the engine size required to attain the goals desired the selection process between a small block (289/302) platform, or large small block (351) platform is relatively easy. So next lets consider engine internals.

Crank, Rod and Piston considerations

To stroke or not to stoke that is the question? Once the base block is selected we need to consider if the engine will be stroked to a larger capacity. The 289 / 302 Windsor small block has two popular stroker crank options, e.g., when using a 30 thou over bore the 331 and the 347 as final capacities. The 331 arrangement does have an option so the piston pin hole does not intersect with the oil control rings. Although you loose the advantage of the longer rod. Note: As long as the control ring support layer is installed correctly oil control is not an issue for the long rod 331 or 347 stroker kits. The costs are almost always similar so from a performance perspective there are few reasons to give up the extra capacity a 347 offers. In contrast, a standard stroke rebuild that happens to require extra costs due to parts refurbishment will often have minimal cost savings over a stroker kit option.

Given we are restricting this to the Windsor family of engine then ask if the vehicle had the 351W as an option if yes then build a stroker 351. I personally see no point in building a plain 351 Windsor as a stroked 302 to 347 (a 3.4 inch stroker) is a superior package in almost all aspects other than the base strength of the block. For a 351W, I would suggest selecting a 408 CID (e.g., a 4 inch stroker) as in my view it is simply the best street engine to build. Alternatively, a 351W stroked to 393 is a suitable approach to get the extra capacity for potentially a lower overall cost given the use of the existing factory 351 rods and easily sourced plentiful low cost 302W piston options. In the case of the 351W platform there is the 418 (a 4.1 inch stroke) or a 427 (a 4.2 inch stroke). The 418 is as large as I would ever consider for a standard 351W block. For a 418 I would limit the top end to 6000 RPM though low end torque would be plentiful for a light car such as a Mustang. Lastly, the 427 option (a 4.2 inch stroke) is pushing space limitations and can have crank interference problems with a standard sized camshaft so leave this size of engine or larger to the employment of a non factory block.

Image 7: A stroker kit includes all the parts required.

If an after market block is selected then the engine bore can be significantly increased. As a result you could build a small block Windsor with a capacity of 363 (or greater) and still have room for future rebuilds. Given a street engine generally does not require the use of an after market block, such as Dart offers, then the extra cost is harder to justify for the potential reliability benefits offered. Though keep in mind, there will be a point where sourcing any good condition factory block will be so difficult or costly that an after market block will be the default choice. This choice does has a side benefit of avoiding possible registration difficulties over pollution control maintenance simply due to the use of a later model engine or other legal barriers that the installation of a later model engine may impose on vehicle registration conditions.

It is important to know the size (CID) of the engine you have decided to build as the capacity determines the air requirement, among other aspects, of the engine. The formula to calculate the engine Cubic Inch Displacement (CID) is;

Formula to calculate engine displacement in cubic inches.
CIDCubic Inch Displacement
STROKECrank Stroke of the engine in inches
BORECylinder bore of the engine in inches
CYLNumber of Cylinders
CID = BORE * BORE * STROKE * 0.7854 * CYL

For example a 289/302 block with a 3.4 inch stroker crank that has a 30 thou over bore will be calculated as;

CID = 4.03 * 4.03 * 3.4 * 0.7854 * 8
CID = 346.95 (so 347 CID)

The purchase of a complete stroker kit does simplify the internal engine parts selection process though if you want to select the parts individually. What is involved in selecting the parts?

Firstly the crank; if intending to reuse a crank then it is required to be inspected for damage. If in serviceable condition then refurbishing would, at a minimum, involve polishing the journals. Otherwise the surfaces would need to be under ground then polished to size. The machine shop will provide guidance on what is required. If the crank is underground then you need to account for this when purchasing the engine bearings. For example, main journals might be underground by 10 thou and rod journals underground by 5 thou resulting in a 10-5 crank. As such you will require over-sized main bearings by 10 thou and oversize rod bearings by 5 thou. Most commonly the shop will apply the same undersized factor for all journals. Lastly the shop will need to know your required targeted clearances before they polish to size. I suggest you use 1 thou for each inch of the diameter of the journal. For example, a 351W main journal diameter is 3 inches so you target 3 thou main clearance while a 302 would be 2 thou clearance. This guideline applies for rod journal clearances as well.

The selection of an after market crank should firstly target the lowest possible external balance for the budget. For example, common externally balanced weights are 28oz and the later 50oz cranks. Only use a 28oz unless the budget allows an internally balanced crank. The general guideline would be to absolutely go internally balanced for longer stroke applications such as a 408 or larger. It is recommended to also go internally balanced on a 302 format if targeting high rpm applications such as extended 7000 rpm plus usage. Similarly good quality cast cranks (such as offered by SCAT) are perfectly fine for high performance street builds though once your have decided on an internally balanced crank then going to a forged crank should be considered.

Image 8: Set of main bearings.

As far as the bearings are concerned a high performance build is well served with Clevite 77 three layer bearings while a quality two layer bearing such as offered by Kings Bearings works well for street performance applications. There is no advantage to purchasing the race 'R' bearings as the performance 'P' bearings are sufficient. Note: A quality standard volume standard pressure oil pump is all that is needed for a Street performance build. A new Mellings oil pump is a suitable choice. If your running larger clearances than suggested then a high volume standard pressure oil pump becomes an appropriate consideration. Do purchase an aftermarket oil pump drive such as offered by Ford performance or ARP. The drive is thicker and stronger than the factory offering to resist twisting, flexing so being able to handle higher RPM oil feed loads.

If the plan involves the use of the stock ford rods then the number one upgrade to perform is to replace the rod bolts with a high tensile ARP (or equivalent) rod bolt. In my mind this is absolutely required regardless of the level of performance you are intending to build. The machine shop will be able to change the bolts and perform a resize of the rods. Do check the total cost as it maybe cost affective to simply purchase a set of after market rods with quality rod bolts already installed. Lastly, the stock rods are press fit to the pistons pins so a superior option, if the piston being used allows, is to convert to floating rod ends. The modification involves installing a very thin brass bush into the small rod end. It is a better setup to use floating rods ends though it is not a high priority in a street build. Note: Most after market rods are usually fully floating rods ends by default so include this as a factor in deciding to reuse factory rods.

Image 9: Example of an I beam (top) and H beam rod.

Lastly, after market rods can be purchased in either a I beam (which is standard) or a H beam format. The majority of builds I prefer to use a forged I beam rod due to the lower weight. This helps in terms of ease in balancing the rotating assembly as well as the performance of the engine such as RPM gain. However for the long stroker engines such as a 408W, the H beam provides the added strength that provides further insurance for the extra load the rods are exposed to under high RPM applications. In addition the H beam rods usually include the stronger CAP screws rather than rod bolts.

The last item to select to complete the rotating assembly is the pistons. There are several factors involved in this process. These include the material type, the valve reliefs, the compression design, attachment method to the rod end. Firstly, the material and construction of the piston determines the level of abuse the piston can endure. The three main material constructions for pistons are cast, hypereutectic (high silicon cast alloy) and forged. The simple guideline is to use cast for stock rebuilds, hypereutectic for mild performance builds and forged for higher performance builds. More often I will simply use forged pistons for the added reliability they offer. If the build is using a performance adder such as turbo or super charging then a forged piston is highly recommended. Note: a forged piston alloy can vary in composition in which case for street performance engines the higher silicon alloy is often the better choice.

A formal approach for piston construction and material selection involves calculating a load indicator where the higher the load the higher grade material required. This approach is based on calculating the mean piston speed at maximum RPM. Using this method provides a clear indication for the preferred type of piston to select. To calculate the mean piston speed use the following formula;

Formula to calculate mean piston speed at maximum targeted RPM.
MPSPiston Speed - Mean speed as feet per minute
RPMTarget Maximum RPM
STROKECrank Stroke of the engine in inches
MPS = (STROKE × RPM) ÷ 6

For example, for a 289/302 block with a 3.4 inch stroker crank (eg a 347) with a Peak 6500 RPM will be calculated as;

MPS = (3.4 × 6500) ÷ 6
MPS = 3683 ft/min

Once the mean piston speed is calculated then my guide applies;

Less than 3200 ft/min safe to use a cast piston
Between 3200 and 3800 ft/min best to use at least a hypereutectic piston
Greater than 3800 ft/min recommend a forged piston

Piston speed is also a factor in the efficiency of the engine as denoted as Volumetric Efficiency (VE). It is impossible to produce the same VE throughout the RPM range. Up to a point, the higher the mean piston speed the higher the potential VE limit. For example, for a street strip engine the highest overall VE curve is produced for a mean piston speed of between 4500 to 5500 fps. It becomes increasingly difficult to make power above 5000 fps average piston speed. The table below illustrates that as you raise the RPM the piston speed increases though at a point VE starts to decrease. Hence there is a point that you would destroke an engine to reduce piston speed while allowing further increases in maximum RPM limits. Naturally this is for very high strung engines that are beyond the scope of this article. On the other end of the scale, you may find it surprising that, on paper, maximum fuel economy is produced at a piston speed of 2000 fps hence the reason modern engines employ long stroke designs. The Windsor engines we are discussing here have relatively short strokes therefore pistons speeds are generally quite low for the RPM in which they are generally used. This hopefully illustrates that these engines are very suitable for higher RPM applications.

Piston speed determines VE limits.
Piston Speed VE Limit
2500 fps86.5 %
3000 fps94.6 %
4000 fps107.2 %
5000 fps117.5 %
6000 fps125.7 %
7000 fps130.7 %
8000 fps130.9 %
9000 fps125.1 %
9500 fps119.9 %

The final aspect not covered in regards to piston construction and material selection is to consider the environment of use. For example, if the engine is going to be abused, such as being exposed to long periods in the higher RPM range, or there is a high chance of engine tune variability or the tune being incorrect at any stage (especially lean) or has a very high compression for the fuel being used hence the tune is more critical then choose a more tolerant piston. In my experience an engine that exhibits pre-ignition problems, aka Pings, will very quickly destroy itself via piston damage. So it is recommended to use quality pistons that are appropriate for the piston speed and intended use of the engine.

Piston choice is also influenced by the cylinder head chamber design and the camshaft lift profile. The main aspects of cylinder head chamber design are the valve sizes (primary the intake valve), valve angles and chamber size. The diameter of the valves determine the required diameter for the valve reliefs. For example, a stock piston has small diameter shallow valves relieves designed for standard angled heads. Given that exhaust valves are similar in size regardless of the intake valve diameter the intake is the critical aspect for Valve relief diameter. If you use a head with a 2.02 inch intake valve then a stock piston relief will not clear the valve if the Camshaft lift profile and associated valve train causes Piston to Valve (P2V) contact. The clearance I generally recommend is 100 thou side clearance. For example, a 2.02 inch intake valve requires a 2.22 inch diameter valve relief while clearance to the relief base surface targets 60+ thou on intake and 100+ thou on exhaust. In general, after market pistons combined with the camshaft used for the street performance engines have large enough valve relief provided though I recommend clearances are checked and verified as a validation step as part of the build process.

The valve relief position on the piston depends on the style of heads being used. The standard ford head layout uses a parallel valve arrangement at 20 deg angle. While the majority of heads fit into this category, e.g, AFR, Dart and many more, the Trickflow systems (TFS) twisted wedge product changes the angle of the valves so they are not parallel combined with non common shallower angles applied to both intake and exhaust. As a result the intake valve relief valve is located in a different position on the piston though in contrast the exhaust remains aligned with a standard relief. There are pistons manufactured with dual relief patterns so either head layout can be used. Lastly, as a comparison only, the Boss 302 or Ford Cleveland head has a cantered valve arrangement where the valve relief locations are different again.

Image 10: Windsor dual relief piston for standard and twisted wedge heads.

Image 11: Boss 302 and Cleveland cantered valve piston illustrating the reliefs.

The top surface shape of the piston is most commonly a flat top or dished for the larger capacity engines. In the case of a 302 based build I prefer to stay with a flat top piston for a performance street engine while a dome would be used for high compression racing engine, e.g., using E85 fuel and 14:1+ compression. In the case of a 351W based engine using a stroker I would select a dished piston to make sure compression stays reasonable. The added benefit is more surface area to push against which slightly helps performance. The static compression ratio I target for a street performance engine is between 10:1 and 10.7:1 which leaves some safety margin for possible poor quality 98 octane fuel use. In practice a 302 engine is difficult to raise the compression to 10:1 without modifications. It is often necessary to deck the cylinder heads to reduce the chamber size. For example, assume you selected a 58cc TFS 170TW head, to move up to a 10:1+ static compression ratio the heads would need to be decked to a 54cc chamber size to help attain the target. Note: Not all cylinder heads provide the material thickness needed to perform significant level of surface decking.

To calculate the static compression ratio we need to calculate the total volume of the cylinder with the piston at bottom dead center (BDC) and then the area remaining when the piston is at top dead center (TDC). To perform this accurately we need to measure various volumes so we have the uncompressed volume and the compressed volume which forms the ratio. The formula to calculate the cylinder volume is;

Formula to calculate volume of a cylinder.
VOLVolume of a cylinder in CCs
DIADiameter of the cylinder
HEIGHTHeight of the cylinder in Inches
VOL = (HEIGHT x (3.14 × ((DIA/2)^2))) * 16.387064

The components of the calculation needed are;

CLV = Engine cylinder volume - this is calculated in CCs
GKV = Engine cylinder gasket volume - this is calculated in CCs
HDV = Engine cylinder head volume - this is measured or accept manufacturers spec
PRV = Piston relief volume - this is measured or accept manufacturers spec
PSV = Piston side volume - this is estimated as 1 to 2 CC
PDV = Piston below the deck volume - this is calculated in CCs

Once the volumes have been measured and calculated the uncompressed and compressed volumes can be calculated followed by the compression ratio;

Formula to calculate static compression ratio of the cylinder.
Uncompressed Volume = CLV + GRV + HDV + PRV + PSV + PDV
Compressed Volume = GRV + HDV + PRV + PSV + PDV
Compression Ratio = Uncompressed Volume / Compressed Volume

The process of performing the calculation is presented for a 302W engine below. First gather the details of the engine components involved. Then follow the simple steps to calculate the compression ratio.

Engine = 306 Windsor, e.g., 30 thou over bore equals 4.030 with a 3 inch stroke.
Head gasket = Gasket bore is 4.10 with a 40 thou compression thinkness
Head Volume = 54cc (decked head to reduce chamber size)
Piston relief volume = 3cc is manufacturers spec
Piston side volume = 1cc
Piston below the deck volume = measured at 8 thou (could be double)

CLV = (3 x (3.14 × ((4.030/2)^2))) * 16.387064 = 626.76127
GKV = (0.040 x (3.14 × ((4.10/2)^2))) * 16.387064 = 8.64964
HDV = 54
PRV = 3
PSV = 1
PDV = (0.008 x (3.14 × ((4.030/2)^2))) * 16.387064 = 1.67136

Calculate static compression ratio of the cylinder.
Uncompressed Volume = 626.76127 + 8.64964 + 54 + 3 + 1 + 1.67136 = 695.08227
Compressed Volume = 8.64964 + 54 + 3 + 1 + 1.67136 = 68.321
Compression Ratio = 695.08227 / 68.321 = 10.17 : 1

The suggested compression for alloy heads on a 302 Windsor is generally between 10:1 and 10.5:1 while for a 347 I would target 10.5:1 to 10.7:1 and a 408 I would target 10.7:1 and higher. In each case this is for a strong street performance engine as the compression ratio does have a relationship with the camshaft. Hence I am assuming a typical street performance camshaft being employed when suggesting the above range of compression ratios. Lastly I target a quench between 30 thou and 50 thou while 40 thou is my ideal. The quench is depth of the piston below the engine deck plus the head gasket thickness. So in the above example, we have a 40 thou gaskets plus the piston is 8 thou below hence the quench is 48 thou. This is near the maximum quench I would generally accept.

The approach I often follow is to zero deck the block, which means the piston is level with the deck of the block at the top of the piston movement, and use a 40 thou compressed head gasket that results in the quench being right on the money. If the quench is to small, e.g., below 30 thou, there is the risk of the piston hitting the head from expansion of the piston due to heat of combustion plus the lengthening of the rod from heat and lateral stresses found at higher revs. If the quench is to large, e.g.,above 50 thou, then this area can be large enough to allow flame to travel into this tight space promoting the engine to ping. It is important to keep in mind the quench area when figuring out your parts combination and compression targets.

Piston ring selection is the last item of consideration. The first aspect is piston ring thickness. Once a list of potential pistons are determined, to further reduce the set of possible choices I first compare the weight, where the lightest piston is preferred, then the thickness of the rings required by the piston is determined. The thinner the rings the better in terms of reduced drag, improved cylinder bore wear characteristics with the bonus of possibly fast embedding behaviour of the ring set. The second aspect is to choose the type and if it is pre-fit or file fit ring package. There are three main ring types, cast iron only, moly and chrome. I personally always select a ductile iron moly faced top ring package. Cast iron for the secondary rings are perfectly acceptable. A stock build normally uses a cast iron top ring while chrome is reserved for extreme heavy duty use and would not be used on a street engine build.

Image 12: Example of a ring package for a 347 stroker piston. Note the support ring is also shown.

A moly top ring can be easily identified by visually inspecting the contact edge to see multiple layers. The center section is brighter material that reflects more light than the outer two sections. The choice of pre-fit or file fit rings is harder. It is always best to file fit each ring to the bore it will be installed into though it can be a very time consuming process that optionally uses a special ring filing tool. Having said that an inexperienced engine builder will find it easier to use pre-fit rings. In either case, make sure the ring package and pistons are available for the machinist to precisely bore and hone to size the engine. The bore to piston clearance needs to be correctly set up as well as the ring gaps while the correct bore hone pattern and depth is required to suite the top ring material chosen.

A good machinist is a really beneficial resource that reduces the chance of later finding issues during the build process. BTW: Do not forget to have the rotating assembly professionally balanced. This is an important pre-assembly step that needs to occur. Once the base rotating components of the engine have been selected your ready to choose the main components that influence the power potential of the engine.

There are four parts, that need to work together, that have the largest influence on the final performance characteristics of the engine. They are the cylinder heads, the camshaft, intake and to a lesser extent the exhaust headers. These parts are the hardest to select as they need to work together to provide the desired result. The easiest approach is to select a top end kit that includes all or most of these parts preselected by the manufacturer to work together. The most common and hence most popular are the kits offered by Edelbrock and Trick flow systems. For the less experienced the kits offered are a worth consideration though they are limited in choice. I prefer to select all the parts individually. The next few sections will continue to guide you through this process.

Cylinder head selection

Selecting cylinder heads is the next major part in the combination selection process. The Windsor engine is fortunate in that there are many options as far as after market cylinder heads available. Although in the past this was not the case so factory heads needed to be used, today unless there is a particular requirement no factory head should be considered. If you must then the GT40 head is the best option. In general once a factory head has been reconditioned, possibly converted to use unleaded fuel, its costs would not be far from the cost of a far better performing set of aftermarket heads. There have been some impressive claims made from builds employing factory heads though in practice for the majority of us factory heads results in disappointing factory performance.

Note: Unlike GT40 heads the GT40p variant have repositioned spark plugs which can limit exhaust header fitment.

An engine is basically an air pump. Cold air flows in and hot flows out. This pumping action occurs at a rapid and varying rate all depending on the current RPM. The most efficient operating range of the engine is determined primarily by the camshaft while the efficiency of the air movement is primarily determined by heads. Both components, e.g., heads and camshaft, should be considered one complete component as they must work well together for a successful performance build. A standard engine has an operating range between 600 RPM and 4000 RPM while a reasonable operating range for a performance street engine is between 1600 RPM and 6000 RPM. Assuming a stock head is at maximum efficiency at 4000 RPM then its use for an engine designed for 50% higher RPM range would not be a suitable combination. Choosing an appropriate head for the engine needs to relate back to the size of the engine and the maximum RPM required for efficient air flow. These factors are first established after which the airflow required from the heads can be determined.

The standard factory Windsor head is based on 20 degree valve angle design hence the majority of after market heads available also follow the factory design. The extra performance is gained by employing larger ports using a superior layout with larger valves and improved combustion chamber shapes. The fact is over the last 50 years cylinder head design has improved remarkably. Most after market heads largely apply the latest ideas. A few manufacturers offer heads where the valve angles are re-aligned to be a shallow angle and often not parallel in contrast to the factory heads. The Trick Flow System's (TFS) Twisted Wedge (TW) design goal introduced the re-aligned layout design that improves flow around the open valve as the valve opens toward the center line of the head. In addition, this approach allows the port length to be shorter thus improving flow while the shallow valve angle provides additional clearance for higher lift camshafts when using stock pistons. Another approach taken is to raise the ports of the heads. This design choice improves the angle the air enters the cylinder chamber where the short side port radius is improved resulting in port flow improvements. In some cases only the exhaust port is raised as is it not so intrusive for the exhaust header to be installed slightly higher. In contrast a raised intake port may require specialised intake manifolds which are generally outside the realm of a street performance setup. In summary standard port heads provide adequate flow for street oriented performance applications while moving valve location and angles can offer many advantages over the 50 year old design.

Image 13: An alloy TFS TW 205 Windsor cylinder head.

An engine can be considered simply an air pump. The goal is to move as much air as possible in and then out of then engine for each rotation. The engine efficiency is improved when pressure differentials and the associate timing of the air movement is maximised. When considering the core parts, rather then guess, I firstly consider the air required and time available as large determinants for deciding that parts required. The starting point is the analysis around the size of cylinder head intake ports. For example, a 302 based engine with a 6000 RPM efficiency limit results in 5 liters of air being draw into the engine every intake cycle in a time period of 0.01 of a second. The intake port therefore needs to flow for one cylinder in one eighth of that time period. In this case, 0.625 of a liter of air needs to flow in 0.0125 of a second for maximum efficiency, e.g., 100% Volumeric Efficiency (VE), to fill all of the available cylinder capacity under standard air pressure. Note: In practice street engines do not reach 100% efficiency though 100% or greater Volumeric Efficiency (VE) is the ultimate goal.

As indicated, I use the size of the engine combined with the Peak RPM to determine the head size required. Hence the larger the engine capacity the larger the head required to support the flow requirements at a given RPM. If a given head is restricted in flow for the given RPM the performance is affected. Equally, the head can also be to large for a given engine for a given RPM. Do not accept Internet gossip that "Head X is to large for a 302" or "Head Y is to small for a 393" etc. It is the context that is important, for example, the same head can be fine for both a 302 and a 393 depending on their respective Peak RPM targets. It is better to have a cylinder head that is to large than one that is to small for a given application. If your not sure always purchase the larger cylinder head.

To select the head, as stated, stay primarily focused on the intake port simply because filling the cylinder is more difficult than evacuating the hot expanding exhaust gasses. One key factor for engine efficiency is the velocity of the port that is influenced by both the size and length (as well as the design and shape) of the port. We will focus primarily of the size and length which is represented as the Cross Sectional Area (CSA) of the port. This is not to say that the overall shape is not important when it is very important. The efficiency of the port is a one of the determining factors when refining the choice between product offerings.

Image 14: An cylinder head intake port with flow loss points illustrated.

As illustrated in the port diagram as shown in Image 8, the walls of the port provide friction to the port flow. So when considering heads, intakes and exhaust headers the port size and length matters, For example, a long port length requires the port size to be larger than if the port length was shorter to flow the same volume of air in the same time period. As an exercise you can prove this to yourself by getting two straws and cut an inch off the end of one straw and joint the two longer straws together. Now we have two ports of different lengths with the same size (diameter) port. If you suck air from them one at a time you will find the long straw will present greater resistance to the air flow.

The first step in determining a suitable cylinder head is to determine the Cross Sectional Area (CSA) required for the Peak HP RPM. It is important to note that I aim for a maximum port velocity of 300 fps (and no lower than 250 fps). If the velocity is greater than 300 fps then the port will start to backup due to sonic choke slowing down the port. This is referred to as choking the port. Note: The design and shape of the port determines the degree of port choke at high velocities. For example a high port cylinder head provides improved direct flow to the valve so it is capable of higher port velocities with lower choke. In any case, I strongly suggest a maximum of 300 fps velocity for your street performance builds.

To calculate the required Cross Sectional Area (CSA) head port size I use this following formula;

Formula to calculate required Port CSA for given Engine and RPM.
CSACross Sectional Area of the port required
RPMTarget Peak HP RPM
STROKECrank Stroke of the engine in inches
BORECylinder bore of the engine in inches
CSA = ( BORE * BORE * STROKE * 0.7854 )*(RPM/360) / 300

For example the required Cross Sectional Area (CSA) for a 289/302 block with a 3.4 inch stroker crank that has a 30 thou over bore at Peak 6500 RPM will be calculated as;

CSA = ( 4.03 * 4.03 * 3.4 * 0.7854 )*(6500/360) / 300
CSA = 2.610

I have provided below a basic table for various engines sizes at various RPM Peaks the calculated Cross Sectional Area (CSA) required. Select the engine then Peak RPM to see the CSA required.


You will note several factors from the table above. Firstly a stock/mild 408 which Peaks at 5500 RPM needs the same head as a 347 with a HP Peak RPM of 6500 rpm. As I will show later a port with a Cross Sectional Area (CSA) of 2.6 inch^2 is a large head. For example, an AFR 205 head has a average Cross Sectional Area (CSA) of 2.5 inch^2. The average punter would say a AFR 205 is to large for a 347 and perfect for a 408. I have shown that for a hot 7000 Peak RPM 302 the AFR 205 is well matched as well as showing that for a performance 347 the AFR 205 is a well matched head as well while the AFR 205 is to small for any 408 beyond the mildest of builds. The large stoker engines require a lot of air and I would argue that the majority of punters with these engines are well undersized. In fact it is very hard to oversize a cylinder head on a large stroker engine build.

Now we are armed with the required Cross Sectional Area (CSA) for our engine build. So we need to review cylinder head products for the appropriate prospective parts. Choosing the manufacturer is often based on budget, personal preferences or recommendations from others possibly from the Internet. I classify head manufacturers and their products in roughly one of four groups. Group one is the exotica group. This group is for race oriented heads or heads that are more involved in terms of the installation process. Group two are my personal favourite heads offering latest designs that result in excellent performance, use quality components and offer great value in their pricing models. Group three are excellent heads though do not quite fit into my personal preferred brands. I would certainly use the heads in this group and would expect to be happy with the result. They are just not, for one reason or another, in my preferred list. The last group are the least preferred. The products in this range are often copies of other products where the castings are mostly fine, unfortunately the valves and springs are often made of lower grade materials hence not suitable for performance applications. If hardware is upgraded to higher quality items then heads in this lower category can operate satisfactorily in a street oriented environment.

Cylinder heads are marketed with usually the size of the ports in CCs and often provide the bench tested air flow numbers shown at progressively higher lift levels. Unfortunately the port potential does is not directly indicated from either figure thus making it hard to perform a general comparison between manufacturers. The professionals measure the port to determine the smallest cross section as this is the bottle neck of the port then factor in the length as shown as this influences possible turbulence affects of the port. Lastly, the efficiency of the port is measured to define the comparison factors for the selection process. Asking an average guy to measure the internal diameters of cylinder heads and flow each product is not a practical approach. So it is best to approximate the average cross section of the port and use the factory flow figures. Note: Often factory flow figures have been attained using exhaust extensions and large engine bores to enhance the flow. So look carefully at the pressure drop, which should be 28 inches of H2O, and the other factors to determine if you are looking at comparable figures.

To calculate the Average Cross Sectional Area (ACSA) we use the port volume and length to roughly determine the equivalent performance. For example, most factory style heads with the 20 degree valves have a port length of 5 inches, while the Trick flow twisted wedge heads have a port length of 4.75 inches, and most high ports such as the Trickflow High Port head has a 5.25 inch length port. Once you know the volume of the port and the length we can calculate the Average Cross Sectional Area (ACSA).

Formula to calculate required average port CSA for an intake port.
ACSAAverage Cross Sectional Area of the port
VOLVolume of the Intake Port in CCs
LENLength of the port in inches
ACSA = VOL / LEN / 16.39

The formula is an approximation that will get you very closely matching head for the intended purpose. For example the average CSA for a typical 20 deg valve for 195cc intake port for a windsor engine will be calculated as;

ACSA = 195 / 5 / 16.39
ACSA = 2.379

Hence a Trick flow 170 twisted wedge head has an average cross section of 2.184, a AFR 185 head has an average cross section of 2.257, a AFR 165 has an average cross section of 2.013, a Trickflow 190 11R has an average cross section of 2.441 and as shown earlier a AFR 205 has an average cross section of 2.502 in^2. Note: The average cross section calculation illustrates the Trick flow 19011R is closer to the AFR 205 than would first appear on the surface. In practice the Trickflow twisted wedge (TW) products with the reduction in valve shrouding from the revised valve angles aligns both heads to be near equivalent for street builds (though very different camshaft specifications are required to be close to optimal).

The required cross section does not indicate the efficiency of the head port. The port shape has a very large influence on the port speed hence mass volume provided to the engine cylinder. It is common practice to have a head ported so as to improve the head flow. On the surface a ported head will result in superior performance than a non ported head. This is not always true. A head can be ported and on the bench at 28in of pressure show higher flow though when tested on a Dyno it may produce no additional performance for the effort. The reason is simple and has been discussed above. If the port becomes to fast, such as over 300 FPS, the air will backup from sonic choke and hence no extra filling capability provided. This simple formula can help identify when porting can assist or when it has the little advantage or even the potential to reduce the performance of the port.

Formula to calculate port speed / efficiency.
FPSSpeed of the port in feet per second
CFMFlow Volume of air.
CSAAverage Cross Section of the port
FPS = (CFM * 2.4 ) / CSA

The formula verifys that port speed at set lift points of the head. For example, an AFR165 head with a cross section of 2.013 in^2 flows, according to AFR, 251 CFM @ .500 lift. The calculated port speed is as follows;

FPS = (251 * 2.4 ) / 2.013
FPS = 299

In this case the efficient AFR165 port speed is already at the designated port speed limit of 300fps so any porting performed without increasing the average cross section will produce no additional mass volume into the cylinder head. In contrast the Edelbrock 170 RPM head with its cross section of 2.074 in^2 and 240 CFM fort flow as a result has a port speed of 278 FPS. Hence it could benefit, and I would recommend, work to the ports. Alternatively, purchase a version of the head with a larger port such as moving from 170 to a 185 to cater to lower air speed only capabilities. Most manufacturers now offer CNC porting either by default or as an option so other than minor work, usually around the throat, have shown there is often little to no benefit to be found in further porting the cylinder head once 300 fps port speed is attained.

In an effort to simplify the selection process the following table is a general starting guide. To select a set of heads for your build look to the size of the engine, then the application followed by scanning left to right within the performance level to select the head. For example, a mid-mild street 302 will work well with a TFS TW170 head leaving some room to grow. While a higher performance 302 build would be well served with the TW19011R.

StockMild streetStreet stripMainly strip
306Stock,GT40.AFR 165, TW170, AFR 185.TW19011R, AFR205.AFR205, TW20511R.
347AFR 165.TW170, AFR 185, TW19011R.AFR205, TW20511R.AFR 225, TFS 225 HP.
408AFR 185,TW19011RAFR205,TW20511R.AFR 225, TFS 235 HP.TFS 245+ HP

Further there are several aspects to consider when selecting your cylinder head. For example, the springs and retainers and the rocker stud size. High spring loads require larger rocker studs. Hence they are available in different sizes usually in 3/8 inch or 7/16 inch. I always select the thicker 7/16 option or purchase a set of studs separately to use simply for the added insurance. Naturally you require a set of suitable rockers for the stud size selected. The spring retainer choice is based on the material used for its construction. Three main materials are available. They are steel, tool steel and titanium. The better the material the lower the weight which assists rocker gear control to potentially allow the use of lower spring rates. Basic steel retainers are fine for the majority of performance street builds though a serious build would use titanium retainers.

The springs need to be capable to handle the lift profile as well as being able to keep the rocker gear on the camshaft for the rev range required. The key is the spring pressures to match the camshaft profile for the weight of the rocker gear. To high a spring pressure can produce high wear on the camshaft and lifters while not enough pressure will allow the lifter to separate from the camshaft resulting in potentially destructive valve bounce. The lighter the valves, springs, rockers and spring retainers the lighter the spring pressure can be and still be able to control the lifter to prevent valve bounce. Quality springs as offered by a number of manufacturers should be used. One such brand is PAC though there are other brands to select from as well.

Cylinder head selection involves matching the engine size, peek intended RPM with the air flow requirements. I have shown a simple process that allows the parts matching selection decision to be refined quickly and easily for a backyard engine builder. Using a number of simple formula this section has shown how to determine the required air mass and defined the cross section required for your engine. In addition, a method has been shown to verify if the port flow of a prospective head is efficient enough hence capable of providing the air mass required at the desired maximum port speed of 300 fps. Lastly, an easy lookup table has been provided for a selection of heads that will work well over a selected range of performance requirements for certain common engine capacities based on the mathematical methods presented.

Camshaft selection

The camshaft controls the heart beat of the engine. Assuming all other parts are working harmoniously together, the camshaft controls the performance range and engine behaviour. Camshaft selection must be matched to other parts especially the heads. Hence, it is the most daunting component to select with confidence. The difficulty is mainly due to the complex nature of an engine and it's operational environment. In addition, there are various external factors that also influence the result. The process detailed below is an attempt to provide a simplified method. Naturally, professional engine builders can have an advantage at selecting camshafts through personal experience, having access to known successful combinations and possibly sophisticated software tools. The average guy does not have access to such a knowledge base. Not all is lost as I will provide a general guideline that if followed accurately will allow a non expert to apply the methods to select camshafts that will work very well as a base to attaining their targeted performance goals.

Image 15: Recommend to use a full size steel core hydraulic roller camshaft with a 351W/HO firing order.

Camshaft selection starts initially with the decision to select the type of camshaft and lifter combination to be used. The early Windsor engine widely employed a flat tappet hydraulic camshaft while a few specialist performance Windsor engines in the early years had a solid flat tappet camshaft. In comparison the later engines moved to a superior hydraulic roller camshaft. These are your main choices though in the after market world a solid roller camshaft lifter design is available. They are usually applied to very high performance engines only. Camshaft design, lifter construction, spring technology, materials and oil technologies have advanced to the point that for a performance street engine I would personally not use anything other than a hydraulic roller camshaft. Very mild street engine would be the only case where a flat tappet hydraulic cam would be considered.

The advantages of the hydraulic roller camshaft includes easier maintenance over what is necessary for a solid lifter camshaft and no requirement for a formal camshaft initial brake in procedure to be followed on first start up as required by a flat tappet camshaft. In addition, a roller lifter camshaft allows recent cam profiles to be used that have fast, in same cases, aggressive lift rates combined with modest open durations to enable large area under the lift curve that allows large volumes of air into the engine. The end result is great streetable performance with the reduced likely hood of over caming the engine. In simple terms the advantages allow superior engine performance to be achieved while providing easy maintenance all while retaining very reasonable fuel economy and street manors.

Image 16: Roller cam provides greater opportunity for air flow due to greater area under the curve for the same total duration.

Flat tappet lifters spin inside the lifter bore of the engine during operation while a roller lifter must not turn at all. So later blocks included a mechanism to maintain the alignment of the roller lifters. This is commonly called a spider. Early blocks to not have support for the factory spider. Fortunately this is not an issue for either early or later blocks as after market linked bar lifters combined with a full sized camshaft is the preferred solution. In general, I would employ linked bar lifters in all of my engine builds even in the factory roller blocks as linked bar lifters are superior, allow greater lift, than the factory setup, while the lifter has greater security from turning than when using the factory roller lifter spider support mechanism.

Image 17: Linked bar hydraulic roller lifters are the best choice.

For the record, an option for early blocks is to install a retro roller camshaft kit combined with a factory lifter support spider by appropriately drilling and tapping the block. These kits contain a small circle base camshaft to lower the roller lifter inside the early block's shorter lifter bore. This can work in mild engine applications though I personally would never install one. The two main issues with an undersized camshaft are excessive camshaft deflection and limited cam lobe profile shape choices. When the camshaft flexes under load it affects the cam profile as presented to the valve as well as total lift provided. Obviously this means the cam lobe selected is not actually operating as required. So as a result the performance is very likely to suffer. In addition, the reduced cam lobe material available limits the options to produce a wide range of profiles. This means there can be considerable difficulty in grinding certain camshaft lobes accurately into the camshaft core. In practice only limited lobes are available to be ground onto a reduced based circle camshaft. Often this means profile that are limited lift in nature with mild ramp rates. For these reasons alone I would not recommend a reduced circle base (style retro fit) camshaft.

Image 18: Retro fit (undersized) camshafts are a bad choice due to inaccurate lobes and excessive deflection under high spring loads .

The material a camshaft is manufactured from is an important aspect influencing the success of a build. There are basically two main materials used, firstly a cast iron core material (that can be hardened) and secondly a steel built. The majority of flat tappet camshafts are cast iron while the majority of factory roller camshaft are based on a steel billet core. In the after market space often roller camshafts are also processed cast iron called Selectively Austempered Ductile Iron (SADI) material rather than a steel billet core. The steel billet is the preferred material to use as it is more durable, has minimal deflection and with its superior strength enables greater spring loads without damaging the cam to control the valve gear at higher revs. I have seen many instances where just changing the cam material to steel and using slightly higher spring rates provides 500 plus RPM to the top end performance curve.

In many cases a special order from your camshaft provider maybe necessary to be supplied with a steel billet camshaft. For example, CompCam has a default SADI core for its camshafts that is indicated via the '-8' at the end of the part number while a steel billet has a '-9'. They fortunately make it easy to order special options that may include producing a catalog cam as a steel core and optionally, as well as the core material used, further customisations such as selecting the lobe family, specific lobe centers and advance offset. This service makes it easy to have your own custom camshaft produced on an as needed basis.

It is relatively easy to identify whether a camshaft is a SADI core or Steel billet. The first method is by holding up one end of the camshaft with a wire and then using a tool, such as a hammer, and carefully tap the bottom of the camshaft. This creates a sound emitting resonance where the steel camshaft produces a higher pitch and persists longer than the iron camshaft. The second method is via inspection. In general a SADI core is only machined to create the lobes and bearing surfaces so looking between the lobes it is easy to identify the dark rough areas that show it is clearly a cast core while a steel billet is completely machined. It is necessary to know what the core material is from a selection perspective as well as from a support parts perspective. For example, the distributor gear and the camshaft retaining plate (generally) need to be the same material as the camshaft so wear properties are even on the surfaces that touch. While my preference is to always use a steel billet core, a SADI core can be acceptable when the lobes are not to aggressive hence the build must be mild.

Image 19: Cast SADI camshaft compared to a Steel billet camshaft.

Lastly, there are two firing orders that have been used over the period of the Windsor engine series. The early firing order was 1-5-4-2-6-3-7-8 and the 351w / High Output 302 is 1-3-7-2-6-5-4-8. On deciding on the cam I will always preference a 351/HO firing order. The reason for the preference is the firing order places the double load on the rear main bearing via the 4 then 8 cylinder double ignition event instead of front bearing for a cylinder 1 then 5 double ignition event. Quite simply, the 351/HO firing order places the load on a stronger section of the engine. Is it a priority? I think it is important and so did Ford as the potentially heavily loaded engines used the later firing order. Having said that, it is important to note that early engines do not seem to mysteriously die because of the early firing order though it was changed at the time the 302 block was cast thinner and a performance version of the engine became available. So for me thats more than enough reason so for that part of my camshaft selection process is to preference the 351/HO firing order.

The basics are now covered so how do we select the appropriately sized camshaft. The first option is to personally look through camshaft catalogs for an off the shelf camshaft that will suite. Alternatively have a custom camshaft selected by a professional on your behalf based on the data you have provided. There are different approaches and levels in custom camshaft services. The most basic approach is the professional selects a camshaft from one of the many providers available, the next level is a camshaft being ground using lobes selected from a camshaft master lobe catalog using a custom Lobe Separation Angle (LSA) and offset, the next is the camshaft lobe is designed specifically for your engine and the camshaft is ground just for you. In theory a custom camshaft is the best approach to attain the performance goals while being matched to the parts being employed. The extra cost for a true custom camshaft needs to be justified against benefit gained. Certainly as your performance goals increase, a camshaft that is closer to optimal is preferred, hence a custom ground camshaft is worth consideration. Other than farming the selection process to a professional, a process is needed to refine the criteria to allow an appropriate camshaft to be identified.

The selection of your own camshaft first starts with understanding the basic terms and what it all means. A diagram best illustrates how the physical lobes and their arrangement relates to the camshaft specification (often called a cam card). In simple terms as the engine rotates so does the camshaft. Assume the lifter is sitting on the base circle of the cam so the valve is fully closed, as the cam turns the lifter will move over the opening ramp. This is the first major event and it documented on the cam card as an Opening Event (OV). This is specified as a degree point relative to piston position at the very top or the very bottom of the engine stroke. As the camshaft keeps rotating the lifter is raised higher to a point where the lift is at its maximum. The maximum lobe lift is also a documented specification on the cam card. Last event of primary importance is the point that the lift reaches the base circle after the closing ramp so the valve is once again closed. Since there are two lobes, an intake and exhaust lobe, for a minimalist camshaft specification there will be at least the four cam events plus the two lift figures.

Image 20: Camshaft diagram to illustrate basic terms such as Overlap, Lift and Lobe Separation Angle.

The cam card will often include a number of additional figures many of which can be calculated from the cam timing events. The extra figures include the duration the valve is off the base circle (usually at 6 thou lift), the lobe separation angle (LSA), the offset or Intake Center Line (ICL), as well as duration figure at set lift heights such as at 50 thou lift (which is the industry standard) and possibly 200 thou lift. The overlap is generally not included on a cam card though it can also be calculated from the cam timing events. In many cases if the cam card does not provide enough information you can look up the lobes in the master cam lobe catalog to investigate any further specification details. Alternatively you can place the camshaft in a V block and use a dial indicator to determine the events and lift profile of the lobe while you rotate the cam though 360 degrees. In practice the cam card would be sufficient to gain enough information to compare for your requirements. So lets review a cam card.

Image 20: Camshaft card specification sheet designed for a 302W.

The cam card tells us it is a hydraulic roller camshaft because the valve adjustment is zero. In the case of a solid roller camshaft, a lash figure would be shown instead. The lobe lift is shown on the next line and following is the valve lift if a 1.6 ratio rocker was used. If it was decided to use a 1.7 rocker on the intake lobe then the valve lift would be 1.7 x lobe lift, e.g., 1.7 * 0.3340 = 0.5678 of an inch which is also conveniently shown in this cam card. The total duration is shown at 0.006 thou lift (the industry standard) and is 275 degs on the intake, while arguably the more important duration figure is shown at 50 thou lift is 219 deg duration on the intake. It is a split camshaft as the exhaust has a greater duration at 50 thou of 222 deg which is a difference of only 3 degs so quite a modest split. The lobe separation is 110 degs. Further the cam card illustrates the actual timing events for two different offsets. Firstly a Intake Center Line (ICL) of 108 which is an offset of 2 degs advanced and another offset of 4 degs advanced. As you view the timing events for the two different offsets the events only move by the change in offset. Normally a cam has the preferred offset ground into the cam shaft so it is installed at that offset if lined up on the standard factory location (often referred to as installed dot-to-dot). Note: for the best result, when the camshaft is installed you need to verify that the camshaft timing events do occur at the points the cam card specifies or adjust (advance or retard) as is appropriate so they do. Lastly, the cam card does not include the overlap period which is not unusual. It is the period from when the intake opens and the exhaust closes which can be easily calculated. In this case at 50 thou lift the Intake valve opens at 1.5 deg BTDC and exhaust valve closes at -1 deg ATDC so add the two figures up gives you (1.5 + -1) = 0.5 degs of overlap at 50 thou lift. Using total duration the overlap is 53 degrees which places it into a mid-range street performance camshaft group. The cam card indicates the camshaft should be an (excellent) mid-range street performance 302W street camshaft primarily via the overlap, good lift combined with expected mid-range (5800) RPM peak durations.

Now that the specification method is understood the process to determine the core camshaft figures from the combination of parts and goals can commence. To this point in the process it may still appear daunting though there are a couple of simplified approaches that I have used to help refine the process. There are many methods some more involved than others so when a process is detailed, as will be further on, often differences in opinion or approach can attract negative comments. If you ask two different cam companies to provide suggestions for the same engine then you will get two different answers. The simplified approach detailed here is designed to get a specification very close to an ideal starting point for the range of street performance engines being discussed. It is not the only approach though it has worked very well for many builds.

The first step in the process is to determine the desired overlap suitable for the goal or targeted use of the engine. Camshaft overlap has one of the greatest influences on the characteristics of the engine. The idle speed, roughness of the idle, vacuum at idle, and potential VE of the engine is all influenced. The choice is a balancing act, for example, high overlap contributes to that tough manly old school idle shaking the car at the lights, the excess overlap reduces the vacuum that may affect the brake booster efficiency, while higher overlap improves the power and efficiency at higher RPM though targeting lower RPM is useful for street driving and cruising. Larger engines by their nature of requiring more air can handle more overlap than a comparatively smaller engine. It is a balancing act between the various behaviours. So being realistic you need to select the appropriate overlap for the size and purpose of the engine. To assist I have produced a set of tables from which to select your target overlap.

Overlap table for 302 CID engine.
Overlap rangeDescription
2549Ordinary Street
4162Street Performance
5874Street / Strip
7895Pro Race
Overlap table for 347/351 CID engine.
Overlap rangeDescription
2651Ordinary Street
4364Street Performance
6077Street / Strip
8198Pro Race
Overlap table for 408 CID engine.
Overlap rangeDescription
2754Ordinary Street
4568Street Performance
6381Street / Strip
86104Pro Race

Find the appropriate table for your engine size and then select the purpose of the engine followed by consideration of the overlap range to determine the target overlap required. If your conservative or inexperienced then choose a value biased toward the start of the range while if you willing to push the boundaries the higher end of the range is appropriate. If unsure then select in the middle of the range. Note: The high end of the overlap range for the street / strip category will be a rough idling low vacuum old school mid to high revs power house if matched up with the right heads and support parts. It is a superior approach to select better heads combined with a milder (overlap and/or duration) camshaft especially if your not experienced or confident.

To illustrate the cam selection steps, lets follow the process for a strong performance street 347 with with a targeted 6000 RPM peak HP. In this case I would review the 347/351 table look at the middle of the "Street Performance" category and choose a 50-55 deg overlap target camshaft. Easy done. Now that the overlap has been decided the next step is to work though the process to determine the total duration. There are many approaches which are often considered propitiatory. The best approach is to base the selection on engine's required intake and exhaust mass. This can become mathematically quite involved. So I am presenting a similar though simplified approach starting with the intake side as this is the most critical of the two lobes. The first step is to calculate the intake airflow, aka, mass, required then select an appropriate Lobe Separation Angle (LSA) from the resulting intake mass required for the target RPM.

Formula to calculate CFM required at Peak RPM.
CPRAir required CFM at chosen RPM
CIDSize of the engine in cubic inches
RPMPeak RPM of the engine
CPR = ((CID / 2) * RPM) / 1728

The formula calculates the air required by the engine at the peak RPM. For example, a 347 at 6000 RPM, it is calculated as follows;

CPR = ((347 / 2) * 6000) / 1728
CPR = 602 cfm

Given an engine is simply an air pump then the air requirements vary depending on both the capacity of the engine and the RPM of the engine. Once we have an idea of engine requirement then refer to a simple table I produced to map the CFM requirement to the appropriate Lobe Separation Angle (LSA). There are few assumptions to this guide; compression ratio is approximately 10.5:1, engine size is between 300 and 450 CID and this is for a performance engine with an appropriately selected set of heads aimed toward supporting the target RPM.

LSA table for Air required.

Using our example 347 the process results in a Lobe Separation Angle (LSA) selection target of 111 degrees. If the compression is lower than 10.5:1 or if the heads are more restrictive relative to the size of the engine then you can tighten the LSA. For example, a large engine (such as a 427) would usually run a little better with a lighter LSA. This is due to the valve area to engine size becoming quite a restriction point. The guide is to adjust the LSA by 0.6 for each point of compression difference. For example, higher compression the LSA is widened or if it is lower than 10.5:1 you tighten the LSA. In the case that low level lift flow of the heads is very good then less overlap and widening the LSA is appropriate. In general the selected LSA is a great starting point for the cam selection process. Now that the camshaft overlap has been chosen and the LSA has been determined we can calculate the target intake duration.

Formula to calculate target intake lobe duration.
DURIntake duration at 6 thou lift in degrees
OVLOverlap for the camshaft
LSALobe Center Angle for the camshaft
DUR = ROUND(2*(LSA + (OVL / 2)))

For example calculating the target intake duration for our sample 347 we are using 52 degs overlap and 111 LSA;

DUR = ROUND(2*(111 + (52 / 2)))
CPR = 274 degs @ 6 thou lift

The intake duration is now determined so the next step is to calculate the exhaust duration. The exhaust volume is expelled via remaining exhaust gas pressure and by the piston movement. As a result the valve size required is always far smaller than the intake. The mass is under pressure though it is still required to be expelled completely just before the exhaust valve closes and no earlier. The intake's air and fuel mass is obviously related to the exhaust mass produced with consideration of the expansion caused by combustion. Hence there is a relationship between the intake port flow that allows the mass into the cylinder and exhaust flow needed to expel the hot exhaust gasses. To represent the relationship the port flow ratio is calculated to determine if the exhaust port flow is sufficient to expel the gasses in the same duration as the intake used to collect the mass in the first place. In determining the ratio my preference is to use the cylinder head flow figures at 300 thou valve lift. The cylinder head's published figures can be used otherwise the head can be tested for the precise actual flow figures. For example, in this exercise we have selected the TFS 11R 190 street port heads which have published figures of intake flow of 205 CFM @ 300 thou and flow the exhaust at 150 CFM at 300 thou.

Formula to calculate the intake port flow to exhaust port flow ratio.
RATRatio of the intake flow to exhaust flow at 300 thou valve lift
IFLIntake flow in CFM at 300 thou valve lift
EFLExhaust flow in CFM at 300 thou valve lift

For the TFS 11R 190 street port CNC heads selected for our sample 347;

RAT = 150 / 205
RAT = 0.73

Balanced flow is considered to be when the exhaust flows 80% of the intake. So a ratio that is lower requires more duration to excavate the exhaust while a higher ratio implies less duration is preferred or is necessary. The calculated ratio for the heads in this example is 0.73 where any figure below 0.80 (80% flow) requires more duration on the exhaust side of the cam lobe while near an 0.80 ratio a single pattern cam, e.g., the same duration as the intake, is appropriate. Lastly, on the odd chance of the flow ratio being higher than 0.80, a reverse split cam is appropriate. In this case the flow ratio difference between 73 and 80 is 7. This is then scaled by 20% then rounded to calculate the split difference which is then added to the intake duration to arrive at the final exhaust duration figure. For example;

Formula for the exhaust lobe camshaft total duration calculation.
EXHExhaust camshaft duration target
RATRatio of the intake flow to exhaust flow at 300 thou valve lift
INTIntake duration in degrees
EXH = ROUND(INT + ((0.80 - RAT) * 120))

This approach assumes the supporting exhaust does not provide any restrictions beyond the port. In many cases a correctly sized exhaust header is not possible or practical in which case a longer exhaust duration could be required even if the cylinder head port ratio was at the balanced point of 80%. In this example the assumption is there that no further restriction exists. So for the TFS 11R 190 street port CNC heads the exhaust duration is calculated with a port flow ratio of 0.73 and a previously determined required intake duration of 274 degrees is;

EXH = ROUND(274 + ((0.80 - 0.73) * 120))
EXH = 282 degs @ 6 thou lift

The main duration figures are now finalised though the LSA offset has not been calculated as yet. In the case of choosing an "Off The Shelf" (OTS) catalog camshaft the LSA offset is generally standardised at a ground in 4 degrees advanced. If considering a custom camshaft then the option is available to nominate an offset. In which case, my goal when calculating the LSA offset is to center the overlap around Top Dead Center (TDC). This is done by calculating the split figure and dividing it by 4 and rounding the result.

Formula for camshaft LSA offset.
OFSCamshaft LSA offset
EXTExhaust duration in degrees
INTIntake duration in degrees

For example, continuing with our 347 camshaft selection example, the LSA offset is calculated with the previously determined intake duration of 274 degrees and the exhaust duration of 282 degrees.

OFS = ROUND((282 - 274)/4)
OFS = 2 degrees advance offset

The calculated advance can be attained at the installation time of the camshaft rather than having it ground into the camshaft. Though you would need a multiple key way timing chain.

The last important figure to estimate is the total intake valve lift. The goal is to lift the valve to a level that is above the point that the valve no longer restricts flow. There are a number of complex formula that can be applied to calculate this with the various valve curtain areas and consideration of valve seat angles and lift. Alternatively, to simplify the process there is a simple guide to use based on the valve diameter.

Formula for camshaft intake valve lift estimate for a roller camshaft.
LFTEstimated required intake valve lift in inchs
DIAIntake valve diameter in inches
LFT = DIA * 0.28

For example, the TFS 11R 190 street port heads have an intake valve diameter of 2.055 inches;

LFT = 2.055 * 0.28
LFT = 0.576 inches valve lift

The calculated target valve lift is 0.576 inch for the intake valve while we also use this as the target for the exhaust. In practice the exhaust lift is more a factor of the lobe family and is usually higher than the intake simply because the lobe profile is the same so with the extra duration it results in a higher lift figure. One option is to use a higher ratio intake rocker than the exhaust though often it is not worth the effort unless your fine tuning the set up or dealing with under sized lift requirements. So we have now completed working out our target figures for a 347 with a peak target of 6000 RPM using TFS 11R 190 street port CNC cylinder heads. The figures are;

Target camshaft specifications.
274Intake duration
282Exhaust duration
111Lobe separation angle
2Degrees LSA offset
0.576Intake and exhaust lift
52Degrees overlap

These are a set of target values as we review OTS camshaft catalogs to determine a set of possible candidates that might match. Focus on the duration figures especially the intake as a priority. Once the initial selection process has completed then perform a second and final review of the candidates to determine the best fitting camshaft to be selected. Note: Do not be tempted to choose a candidate with a larger intake duration figure. Personally I tend to restrict the search to modern lobe designs that have a proven record. For this example I would select a CompCam XE274HR (part no 35-518-8). The specification is 274/282 on a 112 LSA with a 4 degree offset with 54 degrees overlap. The intake lobe lift is 0.348 inches so using a standard 1.6:1 ratio rocker there is 0.555 inches valve lift (which is close enough) while the exhaust has 0.354 inch lobe lift so with a 1.6 ratio rocker the valve lift is 0.565 inches. The intake lift can be made closer to the target lift by using a 1.65:1 ratio rocker which will result in an intake valve lift of 0.574 inches. The exhaust lift would be perfectly fine using the standard 1.6:1 ratio rocker.

Further improvements can be found by requesting CompCam to grind the camshaft onto a steel billet and at the same time select the LSA and offset. In this case, if compression slightly lower than 10.5:1 then I would consider going to a 110 LSA (also as the CFM requirement placed it closer to 110 than the 111). In addition, the 2 deg offset can be ground in on the highly recommended steel billet camshaft core. The cost is minimal and hopefully I have shown it is worth the effort. The last point is the spring selection. In this case the TFS 11R 190 standard 150 lb seat pressure springs are perfect for controlling the fast XE lobe profile. If you select an AFR head (such as the AFR 185) and your using an XE lobe (or similar) then upgrading the springs from the standard 8017 (130 lb seat pressure) to the 8019 (155 lb seat pressure) springs is highly recommended. Otherwise TFS sell their PAC spring package TFS-16306-16 or alternatively the K-motion K-800 springs sets are great for modern fast ramp hydraulic roller camshafts.

The camshaft has been selected so how do you verify that the camshaft is matched to the cylinder head. It should be if the process has been followed though I like to follow a formal verification to make sure the cylinder flow matches the camshaft's intake duration. This same verification process is also helpful if your looking through a parts collection and trying to build an engine from a few old parts lying around or in the case someone else having suggested a particular camshaft. This process first involves calcuating the CFM requirement for a single cylinder as each cylinder has a lobe controlling the air flow.

Formula for a single cylinder required CFM at RPM.
CRSAir required CFM at chosen RPM for one cylinder
CPRAir required CFM at chosen RPM
CLYNumber of cylinders in the engine.

For example, the air required as CFM at a chosen RPM formula was presented previously [CPR = ((CID / 2) * RPM) / 1728] and for our example 347 @ 6000 RPM the CFM required was 602 cfm. So to calculate for a single cylinder;

CRS = 602 / 8
CRS = 75.25 cfm

The camshaft needs to be able to supply at least 75.25 CFM volume of air mass (or more) during the intake valve open duration period. Since the air volume moved during the lowest lift period is very small, I also use the duration figure as provided at the 50 thou lift point. Continuing with the example, the XE274HR has duration of 224/232 @ 50 thou lift so the intake duration figure to be used is 224 degrees. So now the CFM flow required to supply the air mass in that time period can be calculated.

Formula for required CFM flow over camshaft open duration period.
CTSAir flow as CFM for the time slot given by intake duration.
CDICamshaft intake duration at 50 thou lift
CRSAir required CFM at chosen RPM for one cylinder
CTS = (CRS / (CDI / 720))

For example, applying the formula with the single cylinder CFM of 75.25 and camshaft duration of 224 at 50 thou lift provides the required CFM figure that the cylinder head needs to flow at a reasonably low lift level;

CTS = (75.25 / (224 / 720))
CTS = 242 cfm

The cylinder head selected for this example was the TFS 11R 190 street port CNC head. So we need to review the cylinder head flow table and compare it to the required flow calculated previously. The important factor here is that the flow varies depending on the lift of the valve. The lift point of the matching flow will help determine if the camshaft will be suitable.

Flow table for a TFS 11R 190 Street port head (CFM)
Lift (inches) Intake Flow Exhaust Flow

The first item that needs to be checked is to verify the flow required is below the head flow figure for 500 thou lift. If not then I would look to using a (slightly) longer duration camshaft. Secondly, check the 400 thou flow rate and if the required flow is below that figure the camshaft duration could be lowered or a restricted intake manifold could be used without any significant detriment. The maximum lift to match the flow that I would accept is up to 450 thou lift. In this case, 242 CFM is just below 400 thou lift. I like to then add 100 thou to the determined lift at required flow to estimated the minimum required lift for the camshaft. For example, the 242 CFM flow is approximately at 390 thou lift on the TFS 11R 190 heads. Next step is to add 100 thou which results in 490 thou minimum lift required. Comparing to the cam card the minimum lift required is less than 555 thou lift that the camshaft provides thus confirms with confidence the camshaft will be able to provide the air required for the target 6000 RPM HP peak and higher using the TFS 11R 190 cylinder heads.

In summary, we have discussed camshaft basics, type of camshafts, camshaft material, firing order and a detailed camshaft selection process. In general for a street performance engine it is recommended to choose a hydraulic roller camshaft on a steel billet core using a process based on air mass required for the targeted peak RPM. Camshaft selection is one of the more involved process though by following logical steps it has been shown that you do not require years of experience or expensive software to select a suitable street performance camshaft matched to the intended cylinder heads.

Intake manifold selection

Having selected the cylinder heads and a camshaft the next item to choose is the intake manifold. As previously discussed the valves in the heads in combination with the camshaft control the airflow in and out of the engine. The support elements such as the intake manifold, carburettor, and exhaust headers/system must also support the combination. The areas of concern on the intake side are the fueling system, manifold design, port length, and port flow rate.

The first aspect of consideration is the fueling system. There are two primary choices, firstly an Electronic Fuel Injection System (EFI) or secondly a carburettor (or multiple carburettors). An EFI system is available in two main delivery designs. Firstly, a Throttle Body Injection system (TBI) where the unit is affectively an electonic carburettor and secondly a Port Injection System where the fuel is injected directly onto the back of the intake valves. Since affectively the same discussion is applicable for TBI applies as a carburettor the details will be presented later. The better solution is a Port Injection system which provides a common manifold arrangement where the intake ports merge into a larger plenum area fronted by a Throttle Body (TB).

In the case of a street performance engine all that is needed to consider is the size of the injectors, the size of the Throttle Body (TB) and port lengths. When considering the TB sizing, it only needs to be big enough to provide all the air required with no resistance to the flow. For example, we have calculated previously that at peak RPM our example 347 requires 605 CFM. In this case I would select a TB that flows at least 1000 CFM to account for any air flow resistance attributed to the throttle blades. A carburettor or Throttle Body Injection (TBI) would require a lower flow rated Throttle Body (TB) simply because of the atomisation process due to the pressure drop over a venturi for controlled fuel delivery. The fueling point being positioned at the start of the intake port means air and fuel (called wet flow) needs to move along the entire intake port where fuel can pool on the sides of the intake port. While the superior approach is a Port Injection set up that has only air flowing the majority of the length of the port improving precise fuel delivery to the engine. An example port injection EFI manifold for a performance small block (302/347) ford would be the Holley Systemax (pn 300-72S) or Trickflow System Box R performance manifold (pn TFS-51511008). To calculate the TB size for a port injection EFI manifold employing a single bladed Throttle Body (TB) is;

Formula for size required for a single bladed TB.
TBDThrottle body size diameter in millimeters.
CIDEngine Cubic Inch Displacement
TBD = SQRT((196.3 * CID * RPM) / 67547)

For example, applying the formula for our 347 with a target peak HP at 6000 RPM;

TBD = SQRT((196.3 * 347 * 6000) / 67547)
TBD = 78 mm

In the case of using a carburetor the level of flow resistance needs to be considered within suitable operating pressure limits of the carburetor. For simplicity and because of their common use I am limiting the discussion to a Holley style carburetor. The most important aspect of the carburetor is its ability to atomise the fuel properly in the venturi boosters. If employing to large a carburetor it will be unable to get the air shear to atomise the fuel while being to small the carburetor will restrict air flow affecting the performance potential of the engine. So how do you select the correct size? Traditionally, the CFM rating on these carburetors is calculated with a vacuum pressure drop of 1.5in/Hg. While modern carburetors are able to atomise the fuel perfectly well with far lower vacuum pressure drop signals than 1.5in/Hg. For example, in the case of using an old carburetor that required 1.5in/Hg vacuum pressure drop to atomise the fuel may result in a 650 CFM carburetor while a modern carburetor (such as a Holley HP, Quick Fuel or AED) may only require 1in/Hg (or less) of vacuum pressure drop so we can install around a 700 CFM carburetor for further performance due to lower air flow resistance. A simple formula can calculate the minimum size of a modern Holley style carburetor.

Formula for sizing a modern Holley style of carburettor.
CFMCarburettor flow (minimum) required.
CIDEngine Cubic Inch Displacement
VEVolumetric Efficiency of the engine
CFM = (CID * RPM * VE) / 2820

For example, applying the formula for our 347 with a target peak HP at 6000 RPM and expected VE of 0.95;

CFM = (347 * 6000 * 0.95) / 2820
CFM = 701 cfm

In this case I would round up to a 750cfm carburettor since targeting low air flow resistance is a priority as such the larger carburettor would result in higher top end power than a smaller carburettor. Also, it is the calculated value for 6500 RPM where this engine would be shifted in a (drag) racing senario. Note: A mechanisal secondary will provide greater power while a vaccum secondary in this case would provide flexibility where a smaller machanical secondary carburettor would be more responsive in the lower to mid rev range.

In the case of the most common scenario for the type of engines we are focusing on in this article are that they generally use a single Holley style carburettor of which there are two main categories of manifolds. Restricting ourselves for the moment to a single carburettor install there is the single plane and dual plane manifold. In both of these categories there are different heights, port angles to the heads and plenum sizes. The single plane manifold is generally a higher RPM variant where the runners are short, often the carburettor mounting point is higher for an improved entry angle on the port combined with a large plenum for fuel distribution and signal. The lower rev ranges issues for a street drive scenario derives from the active port disturbs the flow of the next port particularly during the overlap period. This is an advantage provided by a dual plane intake where the plenum is divided into two halves where alternate sides provide the intake charge hence isolating the next intake charge from the previous intake event. This works well at low rev ranges though the nature of the design with two plenum areas, different height runners and port cross over with a lot of port angles and turns resulting in slow flowing restricted port designs.

Image 21: A 302W dual plane intake great for low to mid range RPM.

Image 22: A 302W single plane intake great for mid to high range RPM.

The issue for a performance street engine is that it generally needs a manifold that bridges both areas. An example manifold that tries to bridge this middle range is the Edelbrock RPM Air Gap manifold. The design provide larger ports, higher port entry angles and a partially divided plenum. In a performance scenario more plenum is generally needed hence a tall spacer (such as a 1" tall open spacer) combined with a level of porting to help with the air transitions really make the Edelbrock RPM Air Gap become a viable option. Having said that the manifold must be able to flow the air required by the engine otherwise it will restrict and limit the potential performance on offer.

In selecting your manifold, the first step is to establish the flow required of the manifold. We need the manifold to flow as well if not better than the cylinder head for the required air mass. For example, in our example 347 engine we calculated previously that at 6000 RPM target peak we need 605 CFM. This was then applied to the Camshaft duration to calculate that the head needed to flow 242 CFM to provide enough air over the camshaft open duration. This is the base CFM required though the flow reduces as the port length increases. Hence the intake needs to flow more than the base figure to result in no flow loss occurring due to the manifold being present. I use a simple formula to calculate the manifold flow required at maximum lift point of the camshaft.

Formula for required average flow for an intake manifold.
CFMIntake flow (minimum) at maximum camshaft lift.
CFPHead flow required for camshaft duration.
CFM = CFP * 1.10

For example, applying the formula for our 347 with a target peak HP at 6000 RPM and calculated required flow for the chosen camshaft open duration being 242cfm is;

CFM = 242 * 1.10
CFM = 266 cfm

The above figure is based on flow that is provided by the cylinder head, in this case the TFS 11R 190 street port, at 400 thou lift though we lift the valve, using the selected camshaft (XE274HR), 555 thou at which point the cylindar head is capable of 293cfm. Hence using the above formula we have an upper flow requirement of the intake of 322cfm. The mass of air/fuel required to get into the engine cylinder is dependent on the flow, duration and density. The flow varies as the lift changes so we have calculated a lower and upper range of intake flow required of 266cfm to 322cfm. In this case the minimum lift was determined to be just under 500 thou where the head flow is 288cfm resulting in our prefered target flow required of the intake being 310cfm. The appropriate manifolds selection can now be determined from the flow requirements. I have provided a limited table to assist.

Intake Flow table for many popular windsor intake manifolds
Manifold Description Port Flow Length Inches
Ford Stock 2V144 cfm6.5 "
Edelbrock Performer 302205 cfm6.5 "
Edelbrock RPM 302215 cfm6 "
Edelbrock RPM Airgap 302220 cfm6 "
Weiand Stealth 302220 cfm6 "
Edelbrock 302 Victor Jr.280 cfm5 "
Edelbrock 302 Super Victor320 cfm5 "
Edelbrock RPM Airgap 351/W229 cfm6.5 "
Edelbrock 351/W Victor Jr.320 cfm5.5 "
Edelbrock 351/W Super Victor375 cfm5.5 "
Ford GT-40 EFI220 cfm13 "
Trick Flow R Intake 75mm EFI290 cfm11 "

Based on the intake flow range and calculated target the manifold table shows that for our example 347 either a Edelbrock Victor Jr (preferably mildly ported to balance the flow) or the Edelbrock Super Victor would be a suitable choice. The ideal manifold for the purpose of the engine being a street engine would be the Edelbrock RPM Airgap though it would require porting and a 1 inch open plenum spacer. The cost of porting the Airgap manifold to attain close to 300 CFM flowing ports, although possible, might not be worth the minimal benefit in terms street performance gained when employing a dual plane intake. In practice, a driver would not notice any back of the pants loss of low down torque in a well matched combination using a single plane intake such as the Victor Jr. If your unsure and the car is intended to be mostly street driven and accept lower peak HP then a manifold such as a mildly ported Airgap with an open spacer is the best compromise.

The intake length is a tuning and power matching factor that is mostly ignored by most street engine builders though it is instructive to understand this topic as often you will find yourself seeking more HP and this is an area that can be optimised. The intake length can be calculated for an optimal tuned length relative to the pulse arival time to assist with the efficiency of the intake at the chosen peak RPM. We do have limited choices to intake length when we have a small set of after market manifold to select from. Hence this exercise may have limited application unless you are building an intake using an ITB system where you can vary the length of the trumpets or your building a custom manifold. In any case, it can be interesting to go through the exercise and calculate the optimal intake length. The first step is to calculate the Time between intake periods in the cycle.

Formula for the intake wave pulse distance to travel during intake cycle.
PTDIntake pulse travel distance in inches during the intake cycle.
RPMEngines peak RPM
CDUCamshaft duration @0.006 lift from cam card in degrees
WAVWave speed fps for the air density.
PTD = (WAV * 12) * ((60 / RPM) * 2) * ((720 - CDU) / 720)

For example, applying the formula for our 347 with a target peak HP at 6000 RPM, speed of sound at 70 degrees F at sea level = 1128fps and using the selected XE274HR camshaft intake duration of 274 degrees;

PTD = (1128 * 12) * ((60 / 6000) * 2) * ((720 - 274) / 720)
PTD = 84 inches

The length determine of 84 inches is not a practical length to package in engine bay so we allow the wave the run up and down the length of the intake one or more times. This approach allows the intake port to be shorted though the higher the number of reflections the weaker the influence is on intake efficiency. To calculate the various wave reflections are;

Various formula for the determining intake length for various wave length relections.
WL1Intake length for wave X.
PTDIntake pulse travel distance in inches during the intake cycle.
WL1 = PTD / 2 (Wave length #1 - up and back)
WL2 = PTD / 4 (Wave length #2 - 4 times the length)
WL3 = PTD / 6 (Wave length #3 - 6 times the length)
WL4 = PTD / 8 (Wave length #4 - 8 times the length)
WL5 = PTD / 10 (Wave length #5 - 10 times the length)

For example, applying the formula for the first five wave lengths for our 347 with a target peak HP at 6000 RPM with the wave distance of 84 inches;

WL1 = (84 / 2) = 42 inches
WL2 = (84 / 4) = 21 inches
WL3 = (84 / 6) = 14 inches
WL4 = (84 / 8) = 10 inches
WL5 = (84 / 10) = 8 inches

Referring back to our example engine, we have chosen an TFS TW 11R 190 which has a port length of 4.75 inches and we have chosen a Edelbrock Victor Jnr intake manifold with an average length of 6 inches. Both sections total to a length of 10.75 inches. In this case the 4th wave length at 10 inches (actually 10.4 inches not rounded) would have the most influence. In fact it is optimal at 5850 RPM hence it is a good match. Though keep in mind there are variances in temperature and air density which means the rounded approximation is close enough for our purposes. For the last general comment here is the fact that a single plane manifold has two lengths to the intake port so there is a 500 RPM window for optimum wave pulse length. While a dual plane manifold has a far more complex port arrangement with many different lengths in which case optimum RPM is far wider therefore the additive affect is smaller. As a result even if the flow is the same as the single plane manifold the single plane manifold will still have a higher peak HP. Note: If optimising for torque then use the peak torque RPM instead and use a longer length earlier arriving wave pulse.

Exhaust manifold selection

The exhaust system needs to evacuate the exhaust gasses without adding any restriction since all of the calculations and camshaft selection process is based on the valves controlling the air flow. Some would argue that you require some back pressure which in my view is incorrect. The fact is engines produce the most power when back pressure is minimised while maintaining a high exhaust flow speeds. This has been proven over the ages on many dynos. It is important to run the right sized pipe so the CSA is large enough for minimal back pressure while not so large that the exhaust is fully evacuated well before the overlap period. To optimise the intake system fuel air draw involves the exhaust system providing a vacuum pull during the overlap period at the targeted RPM peak. The best exhaust arrangement for our street performance engines are a 4 into 1 exhaust header. There are variations such as a tri-y headers as well as examples that use stepped pipes. In general we want to keep the costs down, easy to source or build and keep selection process simple hence a 4 into 1 header is generally the best choice.

Image 23: Long tube headers for a small block windsor.

As indicated, similar to the intake wave pulse calculations there are exhaust wave pulses being generated during combustion cycle and as the pulses moves along the pipe a vacuum force can be created where the pipe size changes from the primary to the secondary. In our case going from the primary pipe transitioning into the secondary which in the 4 into 1 case is the collector. The best arrangement involves the length of the primaries being the same and each at the optimum length. In addition, the organisation of the primaries inside the collector should be made so the exhaust pulses go around in a circular pattern where the next cylinder firing is adjacent to the pipe with flowing gasses. This approach maximised the opportunity to build a strong vacuum pull on the next intake stroke. Naturally the calculated length of the pipe primaries and collectors, diameter of the primaries and collector and arrangement inside the collector may not be available from an off the shelf header. The choice then can be made to select one as close as possible or have a custom header built. If your using an off the self that is not optimal then I choose the primary pipe diameter as the priority and get the rest as close as possible. The first step is to calculate the length of the primary header.

Formula for the calculation of the primary header length.
PHLPrimary header length in inches
EVOCamshaft exhaust valve open after BBDC degrees
RPMEngines peak RPM
PHL = (850 * (180 * EVO) / RPM) - 3

For example, applying the formula for our 347 with a target peak HP at 6000 RPM and using the selected XE274HR camshaft with the exhaust closing event at BBDC of 77 degrees;

PHL = (850 * (180 * 77) / 6000) - 3
PHL = 33 inches

The length of the header is now calculated at 33 inchs so we can now calculate the primary tube diameter. The size of one of the engine cylinders and the length of the primary is required;

Formula for the calculation of the primary header diameter.
PHDPrimary header diameter in inches
CIDEngine displacement in Cubic Inches
NCYEngines number of cylinders
PHLPrimary header length in inches
PHD = SQRT(((CID / NCY) * 16.38)/((PHL + 3) * 25)) * 2.1

For example, applying the formula for our 8 cylinder 347 with a the primary length of 33 inches;

PHD = SQRT(((347 / 8) * 16.38)/((33 + 3) * 25)) * 2.1
PHD = 1.866 inches outside diameter
*** Round down to 1 3/4 inch primary tube.

The outside diameter of the primary pipe has been calculated to be 1.866 inches so look up the size in the table below and select the smaller size. For example, 1.866 fits between the table entries for 1.75 and 1.875 hence I round down to the 1.75 table entry to bias the size to optimise the RPM between the peak torque RPM and peak HP RPM hence the primary tube size for our example should be a 1 3/4 inch primary tube. Note: If you use the peak torque RPM in the formula set then you round up when selecting the primary diameter tube.

Pipe size selection table.
Diameter Size rounded
1.251 1/4
1.51 1/2
1.6251 5/8
1.751 3/4
1.8751 7/8
2.1252 1/8
2.252 1/4
2.3752 3/8

The primaries have been selected so we move to calculate the collector secondaries. There are involved methods and formula though the simple approach that approximates the preferred length is;

Formula for the calculation of the collector length.
CLNCollector length in inches
PHLPrimary header length in inches
CLN = PHL * 0.5

For example, applying the formula for our 8 cylinder 347 with a the primary length of 33 inches;

CLN = 33 * 0.5
CLN = 16.5 inches long

In this case the collector is optimally sized at 16 inches though most after market headers are far shorter. I would recommend the collector is extended by a further section of the same diameter pipe before necking down to the main tube size. This approach allows the correct length to be built into the exhaust system to optimise exhaust scavenging and as as result enhanced lower RPM engine torque while broadening of power band. Note: a short length collector will be better suited to a higher RPM optimised application.

Image 24: Extend the off the shelf header collector to be the optimum length as part of the exhaust to enhance peak torque and widen the torque curve.

The diameter of the collector is also a simple formula. The heat of the exhaust gasses has not changed to significantly at the point of entry to the collector hence the size is based on the calculated primary tube diameter.

Formula for the calculation of the collector diameter.
CDMCollector diameter in inches
PHDPrimary header diameter in inches
CDM = PHD * 1.9

For example, applying the formula for our 8 cylinder 347 with a calculated primary tube outside diameter of 1.866 inches;

CLN = 1.75 * 1.9
PHD = 3.325 inches outside diameter
*** Round down to 3.25 inch diameter tube.

The main exhaust system's purpose is to allow the hot gasses to exit the system and sound to be muffled with minimal to no air flow restriction. The intake air and fuel mass is combusted turning the oxygen content into carbon dioxide combined with many other gasses and significant levels of heat. Assuming no temperature change the actual mass only increases in size by approximately 7% due to the heavy carbon based molecules replacing the pure oxygen that was part of the air. The mass is hot due to combustion where the initial petrol exhaust temperature is approximately 1400 deg F. The volume gas law states that volume changes in the gas depends on the change in absolute temperature, e.g, V2 = V1 * (AT2/AT1). The exhaust mass only increases marginally while the volume of hot gasses the exhaust needs to flow will be many times the intake volume. The interesting aspect is that the exhaust temperature cools as it travels down the exhaust pipe so the volume of gasses becomes denser hence less flow is need while still being able to offer little restriction. For example, it is best to locate the mufflers as far to the rear of the car as possible as they are often a source of restricted flow. The degree of restriction depends on the design and construction of the muffler. Normally the exhaust uses a single sized pipe so the formula averages out the size needed.

Formula for the calculation of the exhaust flow required with temperature changes.
EPFExhaust flow required in CFM
ITFIntake air temperature in deg F
ETFExhaust air temperature in deg F
CIDEngine displacement in Cubic Inches
VEVolumetric Efficiency of the engine
AVFAveraging factor for temperature changes (use 0.52)
EPF = ((CID * RPM * VE)/3456) * 1.07 * ((ETF+459.67) / (ITF+459.67)) * AVF

For example, applying the formula for the example 347 with an RPM peak of 6000 RPM and a VE of 90% where intake temperature is 80 deg F and combustion is 1400 def F;

EPF = ((347 * 6000 * 0.9)/3456) * 1.07 * ((1400+459.67) / (80+459.67)) * 0.52
EPF = 1040 CFM
*** for a dual pipe exhaust we halve it, e.g., 520 CFM

Using the temperature differential and volumes required, for our example engine, for a dual exhaust system we need the pipes to flow an overall 520 CFM. Straight tube pipe flows approximately 155 CFM per square inch of cross sectional area. A 2 1/2 inch exhaust tube has an area of 4.43 inches so it flows 155*4.43 = 509 CFM hence a dual 2 1/2 inch exhaust system would be suitable. The table below is provided to help determine the tube size relative to the CFM requirement of the main exhaust system.

Main exhaust pipe size selection table.
Size inches Area Flow CFM
1 1/21.48171
1 5/81.77203
1 3/42.07238
2 1/43.54408
2 1/24.43509
2 3/45.41622
3 1/47.67882
3 1/58.941028

An alternative formula combined with the above table can be used for quick calculations to estimate the required size of the main exhaust system. The formula is based on a horse power estimate (HP);

Formula for the calculation of the exhaust flow based on HP.
EPFExhaust flow required in CFM
EHPEstimated engine horse power
EPF = EHP * 2.2

For example, applying the formula for our 347 with an estimated 460HP;

EPF = 460 * 2.2
PHD = 1012 CFM
*** for a dual pipe exhaust we halve it, e.g., 506 CFM

Using the quick formula we have determined that 506 CFM is required hence it matches a 2 1/2 main exhaust system. The flow assumes a straight pipe. When a pipe is bent the flow drops considerably. For example, test have shown a 100 CFM drop will occur in a 2 1/2 inch pipe when compared to a straight section. Packing of the exhaust means bends are inevitable though when planning the exhaust have the number of bends be minimised, bends occur as far to the rear as possible, and be smooth. The best method is to have the exhaust mandril bent rather than the more restictive press bent approach. A larger exhaust tubing size should be used if significant number of bends are to be present in the exhaust system or press bent system is the only option.

Dual exhaust systems normally have the exhaust pulses separated to either bank of the engine. Employing a cross over pipe can assist the system by splitting the pulse and allowing exhaust to travel down both pipes. This arrangement has the advantage of allowing a smaller main exhaust system to perform closer to a correctly sized exhaust. Another advantage is the resulting small noise reduction that will occur while smoothing out the tones of sound produced. There are two general approaches. The first is to simply install a tube 90 degrees to the main pipes crossing to the opposite exhaust system. This is termed the H Pipe balance tube. A superior approach is to merge the dual pipes into an X shape then back to a dual system. Internally the X is opened so gasses have the option to flow down the pipe of least resistance. A variation of the X pipe is to build a chamber that is similar to a flattened header collector of similar length to the collector then exit to continue the dual system. If the cross over pipe is placed strategically then it can assist scavenge the exhaust. This is similar to the operation of a header collector.

Image 25: An example of a H Pipe balance tube.

Image 26: An example of a X Pipe balance tube.

The location of the balance tube can be optimised to help draw the gasses from the system. The best placement position should be where the exhaust temperature starts to cool. This can be determined by using paint points on the exhaust where under normal operation the paint no longer burns off. This is often the best location for the balance pipe. Do keep in mind the practical aspects, such as allowing easier gearbox removal. Ease of maintenance should be an important factor when locating the balance tube.

The last aspect to consider is muffler selection. There are many muffler designs including differences in layout both externally and internally. The primary purpose of a muffler is to reduce the sound to a reasonable (legal) level. As much as sound reduction is important, mufflers are more commonly selected on the quality of the sound though little consideration is given to an often forgotten aspect that the mufflers need to flow the appropriate volume of exhaust gasses hence not be a point of restriction. Very few exhaust installs include muffler flow as a selection criteria which is a continual surprise. Having said that, obtaining flow figures does have its complications as many muffler manufacturers do not publish flow information. A few manufacturers do while a few publish the tests at 15 inches of H2O, rather than at the industry standard 28 inches of H2O, which results a larger CFM figure than it should be if using a standard testing approach.

As an example, see the following table that illustrates the wide range of flow capabilities of mufflers designed for a 2 1/2 inch entry and exit pipe. This list is only a small sample though all the figures have been established by various sources using a flow bench set at a depression of 28 inches of H2O. Using the example 347, flow was calculated that we require 520 CFM flow for each pipe in the dual pipe exhaust system. Hence for this example, reviewing the list of mufflers the recommended product would be the Pypes Race-Pro muffler.

Muffler 2 1/2 inch pipe flow @28 inches selection table.
Description Flow CFM
Pypes Violator (loud)453 CFM
Pypes Race-Pro453 CFM
SpinTech Pro Street427 CFM
SpinTech Sportsman Street402 CFM
SpinTech Truck/RV391 CFM
Pypes Street-Pro349 CFM
DynoMax Super Turbo344 CFM
Hooker Aero Chamber344 CFM
DynoMax Turbo320 CFM
Flowmaster 40 series273 CFM
Flowmaster 50 series213 CFM

The exhaust system is an important component in the overall engine package. A series of formula have been presented to calculate the required system to result in low back-pressure with minimal restriction. Using the example 347 with peak HP at 6000 RPM, the exhaust selected employs a 1 3/4 primary long tube headers that are 33 inches in length, a collector that is 16.5 inches long using 3 1/4 inch tube necking down to a 2 1/2 inch mandrel bent system going into an X pipe (located just after the gearbox) leading into a set of Pypes Race-pro mufflers followed by a twin 2 1/2 inch system through to the rear of the car. The resulting exhaust will be an excellent addition to support the performance street engine being defined as our running example.

Misc Parts selection

There and many parts that make up an engine build and it would be impossible to detail all of them. Having said that there are a few notable items that deserve a brief discussion.

Major fasteners

All major fasteners should be replaced and new bolts, nuts and studs should be purchased. The fasteners under high tension are critical. The rod bolts should be APR high tensile bolts, the main bolts should be replaced with ARP studs while the head bolts should be replaced with the appropriate ARP bolts or for higher performance builds use studs. In the case of the 289/302 blocks the smaller 7/16 bolts are used in combination with a spacer washer or the bolt is stepped. This is due to many after market cylinder heads often using a 1/2 inch head bolt hole so they can be installed on either 289/302 blocks or 351W blocks. The bolts employed to retain the flywheel (manual) or flex plate (automatic) should also be replaced with high grade APR bolts. One aspect to remember is that high quality bolts, when compared to stock, usually require a higher torque limit for correct installation. For example, a 302 ARP head bolt will require 80 ft/lb of torque after the third and final torque sequence round compared to torquing the head bolts with stock bolts at the final 65 ft/lb rate. Note: A 351W APR head bolt will be 100 ft/lb of torque due to being 1/2 inch bolts. So when building the engine verify the correct torque values for the critical fasteners.

Image 27: Main studs for a small block windsor.

Major gaskets

Head gasket choice can be difficult due to so many options in the market place. For a stock to mild engine the Felpro FEL-9333PT1 composite gasket is a good choice. While mild to street performance build could consider the Mr Gasket 5807G steel composite ultra seal head gasket. The best head gaskets for heavier duty purposes are the Multiple Layer Steel (MLS) gaskets. For example, the Felpro FEL-1133 for a standard block or 4 inch bore after market block while use a FEL-1134 for builds employing an aftermarket block such as the dart block with the larger 4.125 inch bore.

For high boost supercharging or turbo applications on a big bore block the FEL-1134SD4 is the recommended head gasket to use. Note: If the MLS gasket is going on a 8.2 deck aftermarket block, it may be necessary to remove the top back rivet from each head gasket, they may lay on the deck on some blocks, such as the dart block. Also, watch for a gap at the china wall as often MLS gaskets are short and do not quite meet the intake gasket. In which case, use some black RTV silicon to fill the gap during the engine build.

During the block machining process remember to inform your machinist to ensure after the block is decked the surface finish is suitable for the intended head gasket. For example, if the build is a street performance engine that has a cast iron block and aluminum heads, and are using conventional steel/fiber composite head gaskets or expanded graphite head gaskets, the surface finish should ideally be 60 to 80 Ra (360 to 480 Rz). Do not go smoother than 40 Ra (240 Rz) or rougher than 100 Ra (600 Rz) with a composition gasket. In contrast, most aftermarket MLS gaskets can handle surface finishes as rough as 60 to 70 Ra micro inches, though some manufacturers specify a smoother finish of 30 to 50 Ra. Smoother is always better for a MLS gasket.

Image 28: Good deck surface is needed for MLS head gaskets.

The intake gasket need to seal the air intake ports as well as the water jacket cooling ports. A stock engine can use the Felpro FEL-MS93334 head gasket while the Felpro FEL-MS95952 is great for intake side of a GT40 head install. In the case of a street performance build a steel backed intake gasket that matches the intake ports such as the Felpro FEL-1262S3 is a recommended choice for the intake gasket. On the other side of the head the Rimflex RFL-3028 exhaust gaskets are the best choice for headers or if necessary the stock replacement gaskets for iron exhaust manifolds. My approach to gaskets are to purchase a complete quality engine gasket kit such as the 280/302 FEL-KS2328 kit and enhance with a few specialist gaskets where necessary. There will be left over gaskets though it is the least expensive and spares can be helpful at times.

Positive crankcase valve

A positive crankcase valve (PCV) is an often ignored component that is important because it maintains the engine internals under a small amount of vacuum rather than pressure. For example, it helps save your main seals, the vacuum improves performance, in fact you would expect 20hp in a decent performance 347, it is a pollution control mechanism as it provides a path for oil vapour to be removed with out polluting the air we breath and it provides a little extra air that a carburetor may need to maintain a fast idle for larger camshaft engines. A catch can can be installed to prevent the oil vapour from entering the engine and possibly poisoning the mixture. In some cases a stock PCV is not suitable for a performance engine as the idle vacuum is lower that stock so an appropriate after market one is required. If a PVC is not employed then an alternative vacuum source should still be used for the engine. Other possibilities are an exhaust based evac or an electric vacuum pump or a mechanical vacuum pump. In any case, it is far better to run a system that keeps the engine's internals under vacuum.

Image 29: The Positive Crankcase Valve provides many benefits.

Fuel Pump

The fuel pump needs to be able to supply adequate fuel to the induction system at the correct pressure. There are two main choices for a carburetor based street performance engine. The first is a up rated mechanical pump that replaces the standard pump. Alternatively, as required by an EFI engine, install a low pressure electric pump. In either case most carburetors critically require no more than 7 psi fuel line pressure. Generally I would recommend an electric pump with an external regulator. An improvement would be to install a return line from the engine bay mounted fuel regulator. In the case of no return line you need to verify if the pump will operate with a dead head fuel line as many will not recommend that type of installation. An issue that can occur with the mechanical pump is over pressurising at higher RPM hence a dead head regulator is advised in this case. An EFI installation requires a high pressure pump that is often installed inside the fuel tank with some form of baffling as they do not like to run dry. As far as possible products are concerned, the Carter electric pump pn P4600HP, are a good choice as they have a reputation to be reliable and are quiet in operation. Depending on the timing cover used if an electric pump is installed the mechanical pump eccentric does not need to be installed and a mechanical pump cover plate added.

Image 30: Carter rotary vane electric fuel pumps are worth considering.

Intake filter

There many designs and options for air filters. Often the selection is restricted by the available space and arrangement within the engine bay of the vehicle. The goal as always is to adequately filter out particles from the air while not restricting air flow to the engine. A serious racing engine build would involve performing various tests on proposed intake air filters to establish the best in relation to the two categories provided. In the case of street engines the general guideline is to provide as much filter surface area as possible to maximise flow while protecting your engine. The traditional paper filter is a preferred filter for a street performance engine build. The wider and taller the filter is the better. For example, I would target a 14 inch diameter filter and look for 2 inches plus preferably 3 inches in height. The filter chosen now leads to the preferred style of air filter base and lid. A perfectly straight base and lid is far from optimal. It is important to allow the filter to fit under the engine bonnet/hood so it is recommended to use a drop base to allow the extra room for the tall filter. This arrangement also has the advantage of directing the air in a smooth flowing curve as the base is rounded for a smooth transition into the carburetter. The lid also promotes the flow transition and compliments the flow from the drop base to redirect the flow up then around and down into the carburetor hence a curved lid is recommended. There are a number of products available for example the Moroso MOR-65910 is a recommended choice.

Image 31: Curved top cover improves air flow over and into the carburetor.

Often filters with the air feed primarily from above or as well as from the sides are available. There are several issues with the style that makes them less than idea. The first issue is from carburetor backfires that can cause an engine bay fire due to the foam filter burning. If you inspect an engine that has run these filters for some time you will often see the foam has been burned. This restricts flow as well as illustrating the fire risk involved. Secondly, traditional carburetors such as the Holley style have four or more air bleeds just within the intake bowl of the carburetor. These have been designed to see air flow over them so the fuel is aerated in the emulsion circuits correctly. If the air is being drawn from above then the emulsion may not work as expect and affect the fuel atomising in affect influencing the tune. Lastly, a cold air feed to the air filter is a affective approach to add performance. This can occur though ducts, intake scoops or many other designs and layouts.

Stall converter

Vehicles that have an automatic transmission behind their performance engine need to use an appropriate stall converter. The camshaft dictates the RPM level at which the engine will produce it's peak torque, which will in turn dictate the optimum stall speed. If your camshaft has an overlap above the street level category or a intake duration of 220-230 degrees (@ 0.050" lift) or above, a higher stall speed converter should be considered. For a mild performance street set up approximately 1000 RPM over stock is the stall speed to target. In general the stall should be rated at about 500-750 RPM under your engine's peak torque RPM. The stall rating is based on a fixed torque and weight figure. So keep in mind the actual stall produced can vary depending on your vehicle weight and engine torque. A lighter car produces a lower stall speed because the amount of resistance (weight) has been decreased. The diff ratio has an influence on the resistance to movement found by the engine. Keep in mind a more powerful engine lowers stall speed due to increasing engine power that has essentially the same effect as decreasing vehicle weight.

Image 32: Higher stall converter is often necessary.

For the example 347 engine with the estimated peak torque at approximately 4000 RPM would ideally employ a 3500 RPM stall converter. If the intended usage was pure street then a lower stall would be perfectly fine though, depending on the diff ratio, likely a little sluggish on initial acceleration. I would not select a stall lower than 2500 RPM. The larger the camshaft the higher the stall converter required. Lastly, high stall converters generate a lot of extra heat. The installation of an external transmission cooler should be considered mandatory with a higher than stock stall speed converter. Heat is the number one killer of transmissions where it is estimated that the majority of automatic transmissions die because of inadequate cooling.

Differential ratio

High performance street engine move the torque and horse power peeks into the mid and high RPM ranges. As a result the lower RPM range is compromised where the engine can labour and be damaged as a result. The primary way to address this issue is by using a differential ratio gear set that is shorter, aka, ratio is higher. This promotes faster RPM increase rates, higher torque levels being applied to get the vehicle moving and more efficient engine operation in the appropriate power range. For example, if engine 1 has 400 ft/lb of torque with a rear ratio of 3.25:1 and engine 2 has 350 ft/lb of torque with a rear ratio of 4.11:1 at the same RPM which would move the vehicle forward the fastest? Well in this scenario the diff ratio is a torque multiplier so engine 1 is presenting 1400 ft/lb at the wheels, e.g, 3.5 x 400 and engine 2 is presenting 1438 ft/lb, e.g., 4.11 x 350, so engine 2 having less HP at that RPM presents slightly more force at the rear tyre. The higher the engine can rev the shorter the gear ratio can be use hence the vehicle can take advantage of the great mechanical leverage provided. A drag race between a more powerful though low revving big block can be beaten by a less powerful high revving small block largely due to the gearing.

This also illustrates why improvement in your drag racing ET can be found with only a change to a shorter ratio. This applies up to a point. So what is the optimum ratio for drag racing? Using the example 347 engine that has a target peek HP at 6000 RPM, the goal is two fold, firstly when the vehicle changes gear we want the power to be close to peek torque to enable fast RPM recovery, secondly we want the RPM to be just past peek HP as you cross the 1/4 mile marker.

Formula for the calculation of the diff ratio for 1/4 mile racing.
ODROptimum differential ratio for drag racing
RPMRPM to cross the line at - just past peek HP
TDMTire diameter in inches
TROTransmission ratio for gear when crossing the line
MPHMile Per Hour speed when crossing the line
ODR = (RPM x TDM) / (MPH x TRO x 336)

For example, applying the formula for our 347 with an estimated peek 6000 RPM HP (where I will add 500 RPM to be past peek) with a 26 inch driven tyre in top gear of 1:1 crossing the line at 120 MPH;

ODR = (6500 x 26) / (120 x 1 x 336)
ODR = 4.19:1

The resulting 4.19:1 ratio has been calculated so the next step is to find the closest ratio below this figure. For a 9 inch diff it would be a 4.11:1 gear set as it would be for an 8 inch diff. Alternatively, a 4.30:1 diff ratio could be considered if the engine has shown to be able to run with slightly higher revs when passing the end of the 1/4 mile. In this case I would stay with the 4.11:1 ratio as it is a street driven car. For a street driven car with a short diff ratio it would be suggested to use an over drive such as an AOD or T5z transmission. In the case of the AOD over drive being 0.67 ratio, this changes the 4.11:1 diff ratio to be a (4.11 x 0.67) = 2.75 final drive ratio. As a result, with the 26 inch tyres, at 100 kph the cruise RPM is approximately 2200 revs. An excellent balance of performance and cruise for a street performance engine.

Engine combinations

The following section is a limited series of engine parts combinations that have been proven to work well together as a package. The primary focus is on the size of the engine, heads and camshaft selected. There is range of combinations based on multiple providers of parts from an almost standard engine through to street strip performance.

289 / 302 cid engines

Stock to mild build 289/302 early windsor.
Heads:Standard Iron heads. Decked. Flat top pistons 9:1 compression plus
Camshaft:Flat tappet - Lunati 10310100
Misc:All stock parts. Use a rebuild kit such as Federal Mogul MHP174-000.
Note: Power range 1200 - 5000. Suggest a 4 barel carbi of 500 cfm.

Mild build 289/302 windsor using GT40 heads.
Heads:Standard Iron heads. Decked. Flat top pistons 9:1 compression plus
Camshaft:Hydraulic Roller - XE264HR 1.7 rockers
Misc:Replace stock rod bolts with an ARP set. Convert heads to stud rockers.
Note: Power range 1500 - 5200. GT40p are a better head though the plug angle
is trouble that is best avoided. Spring upgrade needed such as TFS-2500100.
Consider minor porting in exhaust port, remove thermator and perform some
bowel blending.

Mild build 289/302 windsor using alloy heads.
Heads:Edelbrock E-Street heads. Decked to 54cc. Springs only for flat tappet cam.
Camshaft:Flat tappet - Lunati Voodoo 61001
Misc:Replace stock rod bolts with an ARP set.
Note: Power range 1000 - 5500. 2.02 intake valve is better unless using stock pistons.
1.7 rockers can be used for a little increase in lift.

Street performance build 289/302 windsor using TW aftermarket heads.
Heads:TFS Twisted wedge 170cc heads. Decked to 54cc. Dual Spring option.
Camshaft:Hydraulic Roller - XE266HR
Misc:Replace stock rod bolts with an ARP set.
Note: Power range 1600 - 5600.
650 Holley on a Edelbrock RPM intake manifold with 1/2 inch open spacer. 1 5/8 headers.

**** This is a Boof signature performance street engine combination.

Street performance build 289/302 windsor using TW aftermarket heads.
Heads:TFS Twisted wedge 170cc heads. Decked to 54cc. Dual Spring option.
Camshaft:Hydraulic Roller - TFS2 with 1.7 rockers on intake and 1.6 rockers on exhaust
Misc:Replace stock rod bolts with an ARP set.
Note: Power range 2000 - 6000.
650+ Holley on a Edelbrock RPM intake manifold with 1/2 inch open spacer. 1 5/8 headers.

Street performance build 289/302 windsor using AFR aftermarket heads.
Heads:AFR 165cc heads.
Camshaft:Hydraulic Roller - FW 280H-10
Misc:Replace stock rod bolts with an ARP set.
Note: Power range 2500 - 6000.
650+ Holley on a Edelbrock Air gap RPM intake manifold with 1/2 inch open spacer.
1 5/8 headers and a 2 1/2 inch system.

Street Strip build 289/302 windsor using TW 11R aftermarket heads.
Heads:TFS TW 11R 190cc street port heads. Decked to 52cc.
Camshaft:Hydraulic Roller - XE276HR with 1.72 rockers
Misc:Replace stock rod bolts with an ARP set.
Note: Power range 2200 - 6200.
750 Holley DP on a Edelbrock Vic Jr intake manifold with 1/2 inch open spacer.
1 5/8 headers and a 2 1/2 inch system.

**** This is a Boof signature high performance street strip engine combination.

347 (+351) / 363 cid engines

Mild build 347 windsor using TFS heads.
Heads:TFS TW 170
Camshaft:Hydraulic Roller - Voodoo 20350710 [61010]
Misc:Suggest 1.7 rockers. ARP rod bolts for stock rods if 351W.
Note: Power range 1600 - 5600. Perform some bowel blending.
Edelbrock RPM manifold and 1 5/8 headers to dual 2 1/5 system.
650+ holley with 1/2 inch open spacer.

Street performance build 347 windsor using ARP heads.
Heads:ARP 185 Renegade heads with 8091 Spring package
Camshaft:Hydraulic Roller - XE274HR
Misc:ARP rod bolts for stock rods if a 351W.
Note: Power range 2200 - 6200.
Edelbrock RPM Airgap manifold and 1 5/8 headers to dual 2 1/5 system.
750 DP QuickFuel with 1 inch open spacer.

**** This is a Boof signature performance street engine combination.

Strong Street performance build 347 windsor using TW 11R heads.
Heads:TFS 11R 190 street port heads
Camshaft:Hydraulic Roller - Lunati Voodoo 20350712 [61012]
Misc:ARP rod bolts for stock rods if a 351W.
Note: Power range 2600 - 6600. Rough idle.
Edelbrock Vic Jr and 1 3/4 headers to dual 2 1/5 system.
750 DP QuickFuel with 1 inch open spacer.

**** This is a Boof signature high performance street engine combination.

Strong Street performance [old school] build 347 windsor using ARP heads.
Heads:ARP 205 Renegade heads with 8091 Spring package
Camshaft:Hydraulic Roller - Lunati Bootlegger pn XXX35232HR
Misc:ARP rod bolts for stock rods if a 351W.
Note: Old school cammy idle with strong mid range. Power range 2600 - 6000.
Ported Edelbrock RPM Airgap manifold and 1 3/4 headers to dual 2 1/5 system.
750 DP QuickFuel with 1 inch open spacer.

Street Strip performance build 347 windsor using TW 11R heads.
Heads:TFS TW 11R 205 competition port heads
Camshaft:Hydraulic Roller - TFS3 camshaft. 1.7 rockets on intake, 1.6 on exhaust.
Misc:ARP rod bolts for stock rods if a 351W.
Note: Wide higher end power range. Power range 3200 - 6800.
Ported Edelbrock Vic Jr manifold and 1 3/4 headers to dual 2 1/5 system.
750 DP QuickFuel with 1 inch open spacer.

393 / 408 cid engines

Street performance build 393 windsor using AFR heads.
Heads:185 AFR heads
Camshaft:Hydraulic Roller - TFS 2
Misc:ARP rod bolts for stock rods if a 393W.
Note: Power range 1500 - 5800. Slight noticeable idle.
Edelbrock RPM and 1 3/4 headers to dual 2 1/5 inch system.
750 DP QuickFuel with 1/2 inch open spacer.

Street performance build 408 windsor using TW 11R heads.
Heads:TFS 11R 190 street port heads
Camshaft:Hydraulic Roller - Lunati Voodoo 20350712 [61012]
Misc:ARP rod bolts for stock rods if a 393W.
Note: Power range 2000 - 6000. Slight noticeable idle.
Edelbrock RPM Air Gap and 1 3/4 headers to dual 2 1/5 inch system.
750 DP QuickFuel with 1/2 inch open spacer.

**** This is a Boof signature performance street engine combination.

Street Strip performance build 408 windsor using TFS Highport heads.
Heads:TFS Highport 235 CNC ported heads
Camshaft:Hydraulic Roller - XFI248HR
Misc:4000+ RPM Staff converter.
Note: Power range 3000 - 6500. Noticeable idle.
Edelbrock Super Vic and 1 7/8 headers to dual 3 inch system.
850 DP QuickFuel with 1 inch open spacer.

The possible parts selection combinations are impossibly large. A number of combinations have been detailed and although will work well are in most cases involve compromises. In particular due to off the shelf camshafts that are based on a SADI core. In general it is recommend to consider a custom camshaft grind based on a steel camshaft core be employed even the most basic performance builds. The use of linked bar lifters are suggested. Lastly always verify the cylinder head springs are suitable for the RPM range and lift intended.


This technical article has covered all of the major components and parts while providing a process to select the right parts for street to street/strip performance engines. The author has found no other article similar anywhere so it is hoped this will prove to be a valuable source of information to many readers. To enhance the article, an example engine parts list is created from a base goal to provide a grounded illustration of the steps and processes. All the while detailed consideration of goals, engine application and component benefits have also been discussed. The processes and methods shown are not the only way to approach the parts selection process though for the purpose described it has proven to work very well in practice. There are always compromises though using a process, such as described, will improve the chances of a complete success of the final build.

The next phase is to build the engine. The author intends to provide an article with the instructions and guides to perform the actual building of the engine from the parts selected. This will included; Block preparation, Engine building in a step by step process, Initial start and camshaft running in process, as well as the embedding in the rings and engine running in techniques. This work will focus once again on the Windsor small block V8 though the process is applicable to most engines. I hope this work is useful and answers the majority of questions. If not a Frequently Asked Questions section is provided and will expand as points of interest need clarification from reader feedback.

The author wishes you well on the wonderful journey of engine parts selection and engine building. The goal is to encourage the reader to "Have a Crack" and enjoy the great satisfaction of building your engine - all the information and more has been provided to assist with the journey. Enjoy.

Cheers Boofhead.

Frequently Asked Questions

This section is to provide answers to common questions that have resulted from feedback from the readers of the article. This is a living section in that questions will be added as needed over time. Do not be afraid to ask questions as this is the only way to learn.

Question:Where have the flow figures come from?
Answer:The flow rates have been published by manufactures or professionals or from open forums.
All tests were at the same pressure drop. It is always best to independently verify the
values as testing techniques and calibration of equipment can vary the results.

Question:The AFR and TFS TW 11R appear to be preferred cylinder heads - why?
Answer:These two manufacturers are now dominating the after market with their respectively
excellent products. They both offer latest port designs that flow very well while offering
quality light 8mm valves combined with excellent PAC springs at a very good price.
Note: Order the 8019 springs with AFR heads.

Question:Why are the standard Windsor iron heads undesirable for a performance build?
Answer:The standard ford Windsor heads are intake port sized at 126 cc with a poor port design.
As such the head has a cross sectional area of 1.538 in^2 which is marginal for a stock 302
as it will be done by 4000-4300 RPM. A head this size in a 347 would be done by 3500-3800 RPM.

Notice: This article is Copyright (c) 2016 to the Author known as Boofhead on Mustangtech.com.au.
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