351 CLEVELAND BASICS AND PERFORMANCE TUNING
FOR HIGH PERFORMANCE-STREET CARS AND SPORTS CARS
During the 1960s Ford had manufactured small high winding V8 performance engines that demanded a lot from the driver and were some of the most expensive engines Ford manufactured. Although such engines provided spirited performance in lighter vehicles such as the Shelby Cobra and the 1965 Mustang, they were not as well suited for powering heavier vehicles. Ford had also manufactured large V8 performance engines that overwhelmed the chassis and tires of its production cars with low rpm torque and made it all too easy for the driver to lose control. The large engines were also detrimental to a vehicle’s handling due to the amount of weight they added to the front of the vehicle. After those two extremes Ford settled on manufacturing a mid-size V8 with a heavy dose of mid-range power. And just like in the story of Goldilocks and the Three Bears, that was just right.
The 351C proved THERE IS A REPLACEMENT FOR DISPLACEMENT!
Link: The Sound of a Charging Rhino
Link: A Charging Rhino at Lake Nagambie
Link: A Charging Rhino at Bathurst
Link: A Charging Rhino at Spa-Francorchamps
Link: A Charging Rhino Revving to 7K on an Australian Back Road
BECOMING FAMILIAR WITH THE 351C
Part 1: Terminology and Production Information
Part 2: Durability
Part 3: State of Tune
TUNING THE 351C
Part 4: The Combustion Process (thermal efficiency)
Part 5: Induction System (volumetric efficiency)
Part 6: Exhaust System
Part 7: Camshaft and Valve Train
RACING WITH 351C POWER
Part 8: Preparing A 351C Racing Engine
OTHER 351C RELATED TOPICS
Link: Unleashing the Performance Capabilities of Cobra Jet (Q Code) Engines
Link: 275/285 Custom Street Cam
Link: Ignition Upgrade - Ford EDIS (Distributorless)
Link: 500 or More Horsepower From Small Block Fords
Link: 408 Cubic Inches Using All Ford Parts
PART 1 - TERMINOLOGY AND PRODUCTION INFORMATION
For those of you new to the 351C or those who have questions about the terminology, a quick overview. The 351C is a member of the engine series Ford named the "335 Series". There are 3 pairings of engines in this series.
In Ford terminology a 351C equipped with a 4 barrel carburetor was referred to as a 351-4V; 4V did not mean 4 valves as it does today, rather it meant the carburetor had 4 venturis. For the same reason a 351C equipped with a 2 barrel carburetor was referred to as a 351-2V. The US manufactured 351-4V was the focal point of the engine series in terms of high performance, it was installed in Ford and Mercury automobiles in the USA and Canada, it was exported to Australia and installed in Australian Ford Falcons and Fairlanes, and it was installed in several sports cars manufactured in Italy and one sports car manufactured in Australia.
The 351-4V was manufactured in 3 versions we refer to using Ford's engine codes for those engines; i.e. the M-code engine (351 4V), the Q-code engine (351 4V GT or 351 4V CJ, i.e the 351 Cobra Jet version), and the R-code engine (351 4V HO, i.e. the 351 Boss version). The 2 barrel carburetor version is referred to as the H-code engine for the same reason. All 3 versions of the 351-4V were tuned for higher rpm with cylinder heads having raised intake ports of larger cross section than the heads found on the 351-2V. Those heads were known as 4V heads because they were designed for engines equipped with four barrel carburetors. The 4V heads were also initially equipped with larger 2.19" intake valves and 1.71" exhaust valves (1970 - 1972). The cylinder heads installed on 351-2V engines were referred to as 2V heads, they were equipped with 2.04” intake valves and 1.65” exhaust valves.
Link: 351C US Production History
PART 2 - DURABILITY
351C SHORT BLOCK DURABILITY
Outside of the 1971/1972 351 4V HO engines (351 Boss) all US manufactured 351 Cleveland short blocks were basically identical with two exceptions: (1) some engine blocks had two bolt main bearing caps, others had 4 bolt main bearing caps; (2) the 1970 through 1972 pistons were flat top pistons, the 1973 and 1974 pistons were dished pistons.
The blocks were all cast from the same type of iron, they had the same thin cylinder walls (nominal 0.160") and they had the same bulkhead thickness. The main bearing caps were all equally robust, the only difference being some were cast for two bolts, others were cast for 4 bolts. Naturally aspirated 351 Clevelands have never required aftermarket steel main bearing caps, splayed bolts, etc. The main bearing caps secured by 2 bolts resisted "walking" until about 8000 rpm! Therefore 4 bolt main bearing caps were not needed for a street-performance engine, a sports car engine, or any engine limited to about 7000 rpm. The measures mechanics take to stabilize the main bearing caps of a small block Ford have never been necessary for a 351 Cleveland. The production block was admired for the sturdiness of its bottom end and it was notorious for its thin cylinder walls and its lubrication system issues.
All 351 Cleveland crankshaft castings were also identical, they were all cast from the same high nodularity iron alloy, they were all externally balanced, and they were all equally strong. The crankshaft was partially counter-weighted (having 6 counter-weights) as opposed to being fully counter-weighted (having 8 counter-weights). This type of crankshaft is common for mass-produced crankshafts because it is less expensive and less difficult to manufacture and results in a lighter crankshaft. It was also designed to maximize bob-weight as opposed to being designed to minimize bearing load. This is common for cast iron crankshafts because cast iron is less dense than steel, therefore cast iron crankshafts usually require external balancing. The goal in maximizing the crankshaft's bob-weight is to minimize the amount of external weight required to balance the crankshaft. There were no drawbacks to this type of crankshaft at the engine speeds for which the engines were designed, but all such crankshafts heavily "load" the second and fourth main bearings and bulkheads during ultra-high rpm operation (>8000 rpm). This didn't stop people from using the crankshaft for racing however! They have been used for 8,500 rpm NASCAR racing and 10,000 rpm Pro-Stock and Super-Stock drag racing. They were pretty tough crankshafts. Unfortunately the blocks took a beating at constant ultra-high rpm usage, the production blocks would eventually develop cracks at the second or fourth main bearing saddles that would extend upward through the bulkheads and crack the walls of the adjacent cylinders. This is not a concern for street performance or sports car engines however.
All 351 Cleveland connecting rods were identical forgings, they were all forged utilizing the same heat treated 1041 steel. The factory rod nuts were a source of trouble, they had a reputation for stripping their threads at high rpm. The factory connecting rods were strong enough however for sustained 7200 rpm racing once they had been properly shot-peened and equipped with 180,000 psi rod bolts; but only the connecting rods found in the 1971/1972 351 4V HO engines were prepared this way by the factory. Likewise only the 1971/1972 351 4V HO engines were equipped with heavy-duty forged pistons and crankshaft dampers that were bonded and heavy enough for engine speeds above 6000 rpm. All other 351 Cleveland short blocks were equipped with cast pistons and unbonded, lightweight or medium-weight crankshaft dampers.
Since the short blocks are all identical in strength it is possible to build an equally durable 351C performance engine no matter which short block you have on hand as a foundation. Round skirt pistons resolve the cylinder wall cracking issue, and tappet bore bushings correct the deficiencies of the lubrication system. The crankshaft is more than durable enough for street applications, and stress cracks at the second or fourth main bearing saddles are not an issue for engines operated at or below 7000 rpm. The factory connecting rods employing pressed wrist pins and equipped with ARP rod nuts (or the complete ARP bolt & nut kit) are plenty strong for a street engine and an occasional blast to 7000 rpm.
When we spend time and money improving the performance of an engine, it is a natural assumption that we are doing so because we plan to take advantage of that increase in performance from time to time. We plan to test the limits of the car's chassis and stress the engine to a greater degree. None of us want the engine to fail while we are "flogging" it. So there are a small number of 351C problem areas to set straight as a precautionary measure, thereby insuring the things that go terribly wrong from time to time don’t happen to your car’s engine. The durability of other parts should be improved to insure the engine can sustain higher output, insure it can sustain operating at higher rpm, and insure it can endure high performance driving (i.e. being flogged) without damage.
(detailed information regarding Manley valves, valve springs, push rods and rocker arm bolts is provided further below in PART 7)
Every way in which the 351C design deviated from the design of the SBF was to make an improvement, with one exception ... the lubrication system. The SBF utilized three lubrication passages; oil was supplied to the crankshaft main bearings first via a dedicated passage, then at the rear of the block the oil supply was split into two additional passages to supply the two banks of tappets; this is referred to as a main priority system. To save money the 351C was designed with only two lubrication passages, one for each bank of tappets. Lubrication for the crankshaft’s 3 central main bearings was supplied by branches intersecting the same oil passage shared by the right hand bank of tappets.
Ford found it necessary to redesign the tappets installed in the 351C due to the large port in the wall of each tappet bore, a result of the way in which the tappet bores intersect the oil passages. Tappets designed for the SBF and 351W allowed too much oil to flow to the 351C valve train. When a 351C equipped with the "wrong" tappets is operated at higher rpm the rocker covers flood with oil while the oil pan is slowly pumped dry at the same time. Thus the 351C has compatibility issues with tappets having certain types of oil metering designs; this also illustrates the importance of limiting the amount of oil flowing to the 351C valve train.
Lubrication system pressure is supposed to be controlled by a relief valve built into the 351C oil pump. The setting of the relief valve is controlled by a spring intended by the designers to maintain 60 psi nominal hot oil pressure (50 psi minimum and 70 psi maximum) but the 351C has a propensity for low oil pressure. Hot oil pressure below 50 psi indicates an excessive amount of oil is flowing into various "leaks and clearances”, overtaxing the capacity of the oil pump. The 351C also has a propensity for bearing wear. The symptoms of insufficient lubrication are evident even in low mileage engines; those symptoms include ribbons of bearing material lying in the bottom of the oil pan, scoring on the bearings, bearings being polished, or bearings worn so much they are no longer silver in color but copper colored. Obviously the excessive amount of oil flowing into various "leaks and clearances" is not flowing to the rod bearings!
The two basic design flaws of the lubrication system are:
(1) There's no control of where the oil is flowing nor is there control of how much oil is flowing.
(2) The large ports in the walls of the tappet bores create three additional problems.
People are under the assumption 351C lubrication problems occur predominantly above 7000 rpm, but cavitation does not turn off and on like a light switch at a specific engine speed. Cavitation increases gradually, becoming more severe as the speed of the tappets increases; cavitation is therefore impacting the oil passages to a lesser degree at engine speeds below 7000 rpm. Cavitation simply increases to the point of causing rod bearing failure at some point beyond 7000 rpm. But any amount of cavitation in the right hand oil passage shall impede the flow of oil to the crankshaft's 3 center main bearings to some degree.
There are those who insist the lubrication system is "good enough" up to 6000 rpm. Yet even low-mileage engines equipped with factory camshafts were plagued by low oil pressure and worn bearings. High lift rate camshafts and typical high mileage tappet bore wear worsen the problems. The lubrication system’s performance also worsens as engine speed increases; the performance diminishes to the point of rod bearing failure at some point beyond 7000 rpm. The bearings for connecting rods #2 through #7 are affected, but the bearings for connecting rod #2 or connecting rod #7 are usually the first to fail. All of the lubrication system problems impact solid tappet motors and hydraulic tappet motors equally. The symptoms are the same regardless if the rev limit is 5000 rpm, 6000 rpm, 7000 rpm or higher; the symptoms merely worsen as rpm increases. The consensus has always been that any 351C being rebuilt for any kind of performance application (from mild to wild) needs some improvement to the lubrication system.
Improvements to the 351C lubrication system should focus upon correcting the design flaws rather than the symptoms. One corrective action would be to modify the lubrication system to better control where oil is flowing and to better control how much oil is flowing. We can both minimize the excessive amount of oil flowing to waste via the tappet clearances and limit the amount of oil flowing to the valve train by installing 16 tappet bore bushings. We can also limit the amount of oil flowing to the camshaft bearings by installing 5 cam bearing oil passage restrictors. If we allow oil to flow unrestricted to the crankshaft after making those modifications we are essentially giving lubrication of the crankshaft priority, i.e. we've succeeded in modifying the 351C lubrication system to behave as a main priority system. Thus modified there is plenty of oil volume even with the standard volume oil pump, the standard oil pump spring shall operate in the middle of its range and control oil pressure at about 60 psi in the manner it was originally intended to do, and the quantity of motor oil flowing to the crankshaft shall be substantially increased at all engine speeds, even low rpm!
The tappet bore bushings also correct the other design flaw; they eliminate the large ports in the walls of the tappet bores, metering oil to the tappets via small orifices instead. This isolates the oil passages from the motion of the tappets thus eliminating cavitation in the oil passages; this is another vital step in making it possible for oil to flow unimpeded to the central 3 main bearings. The tappet bore bushings also resolve tappet compatibility issues.
The reasonably priced do-it-yourself tappet bore bushing installation kit available from Wydendorf Machine (selling for $400 USD) makes all of this affordable and within the budgets of a large range of engine projects. Wydendorf Machine
Four decades ago only eight bushings were installed in the right hand tappet bores, but it is customary these days to install bushings in all sixteen tappet bores. The reasons for this are to insure consistent oil control at all sixteen valves, to perform optimally with hydraulic tappets, and to resolve 351C tappet compatibility issues. The tappet bore bushings remove the task of oil metering from the tappets or push rods, and they make the 351C more tolerant of which type of tappet is installed.
My preference is to drill the bushings with 0.060" (or 1/16") orifices for all hydraulic tappet applications, all street and sports car applications, and all road racing and endurance racing applications because once the bushings are pressed into the block the orifice size can't be changed. The 0.060" orifices are a good size for a general purpose or "do-it-all" type of engine set-up.
Four decades ago four restrictors were installed in the passages supplying oil to cam bearings #2 through #5, but it is customary these days to install restrictors in all five cam bearing oil passages. Experience has proven an oil passage restrictor for cam bearing #1 improves the performance of the lubrication system, therefore it is assumed the oil passage for cam bearing #1 diverts a significant amount of oil from the main oil passage which it intersects. Acquiring five camshaft bearing restrictors shall require purchasing two Moroso #22050 restrictor kits, because each kit only has four cam bearing restrictors. The restrictor for cam bearing #1 is installed in a different manner than the restrictors for the other four cam bearings; it must be installed more deeply within the cam bearing oil passage so that it restricts oil to the #1 cam bearing and not to the #1 main bearing. The large restrictors included in the Moroso kits are not used.
PART 3 - STATE OF TUNE
The 4V version of the 351C was not a low performance engine requiring a bunch of aftermarket parts to turn it into a hot performer; it was equipped with high port, big valve cylinder heads which were tuned for peak horsepower at 6000 rpm. It had outstanding thermal efficiency by virtue of the high turbulence combustion chambers. It was also equipped with a 750 cfm carburetor and a high lift camshaft (in all but the M code version). It was obviously intended to be a high performance engine off the showroom floor. The engine's performance can be described as having good drivability, a strong dose of mid-range power, and the willingness to rev to high rpm. This is an ideal power characteristic for a high performance street car, a sports car, or a GT car. Over the decades I've read reports praising the engines in high end sports & GT cars such as Ferrari, Lamborghini, Mercedes and Aston Martin for having similar power characteristics. This characteristic of the 351C 4V is something I try to avoid diminishing in any aspect when I tune the engine for higher output.
Unfortunately the 351C was manufactured in an era when air pollution standards out-paced automotive pollution control technology, therefore the engine's performance suffered. Ford never manufactured the 351C in a version that realized the engine’s full potential. Ford published a "guide" for hot-rodding the 351C in 1970 (Autolite publication #MP-1046) which I realized by about 1975 was not so much a guide for hot-rodding the engine as it was a guide for de-smogging the engine. The guide did not recommend any special high performance parts, the recommendations centered around production parts. The obvious reason for this, the 351C with 4V cylinder heads was a high performance engine off the show room floor. I consider de-smogging the engine the same as optimizing it to perform in the manner Ford originally intended. Ford rated the output of a 351C modified according to their guidelines, equipped with the GT/Cobra Jet hydraulic tappet camshaft, at 65 horsepower above the 1970 factory 351 4V specification. That's 365 horsepower at the flywheel, or 290 horsepower to the rear wheels! So you see, there's quite a bit of performance built-into the factory engine. Thus equipped the engine operated on premium pump gas, i.e. gasoline rated 91 octane in the US and Canada or rated 95 octane everywhere else. The engine idled well, it had good manifold vacuum and it retained the drivability typical of a factory engine. In fact drivability improved. Even those owners who are opposed to “hot-rodding” the engine in their car would enjoy the improvement in low rpm pep and drivability that can be achieved by simply reversing the smog-tuning of their car’s engine in this way. Ford's 365 horsepower version of the 351C is the baseline for experiencing the performance the 351C truly offers. A 351C in this state of tune may prove to be more than enough for a great number of drivers. Here's a summary of that guide for those of you who are curious.
I'd like to make two observations:
(1) the 1971 351 Cobra Jet engine, rated at 280 horsepower, had a specification similar to this except it was equipped with open chamber D1ZE 4V cylinder heads, 8.7:1 compression ratio, a 750 cfm Autolite 4300D carburetor and cast iron exhaust manifolds. 85 horsepower was lost in the process!
(2) Two parts not mentioned in the guide, but available in 1970, were the D1ZZ-6250-BX hydraulic tappet camshaft and the Shelby intake manifold. The camshaft specs were:
290°/290° advertised duration
219°/219° duration at 0.050"
0.505"/0.505" valve lift
114° lobe separation angle
It would have added 16 horsepower to this combination, stretching the output to 381 horsepower. The Shelby manifold would have added another 20 horsepower. Combined those parts would have raised the potential output of this combination to 401 horsepower. That was a lot of horsepower in 1970, especially from an engine of this displacement. It could be achieved without ultra-high compression, without super-premium gasoline, without a solid tappet camshaft, without dual four barrel carburetors AND without diminishing the engine's drivability.
COMPARING THE CYLINDER HEADS
A comparison of the differences in the state of tune of the 351 2V and the 351 4V boils down to a comparison between the differences in performance and power characteristic imparted by the cylinder heads.
One aspect of cylinder heads that impacts the performance of an engine is their combustion chamber. The combustion chamber volume affects the engine’s compression ratio, and the combustion chamber’s design affects the maximum compression ratio that can be utilized for any given gasoline octane without detonation or pinging. The combustion chamber design also affects the engines thermal efficiency and therefore its horsepower output. The 351C 2V cylinder head and 4V cylinder head share the same poly-angle wedge combustion chamber design, and therefore they share the same excellent thermal efficiency. However, as one would expect, with differences in valve sizes, port heights and port cross-sections the cylinder heads differ in volumetric efficiency. The maximum air flow performance of an unported 2V intake port is about 200 cfm at 0.400” intake valve lift; equivalent to approximately 400 horsepower. The maximum air flow performance of a fully ported 2V intake port is about 250 cfm at 0.500” valve lift. That’s not bad performance, that much air flow is equivalent to approximately 500 horsepower; which is more than enough horsepower for any vehicle equipped with street tires.
In absolute terms the 4V cylinder head is a better cylinder head for high performance because it has higher volumetric efficiency. The air flow performance of an un-ported 4V intake port is 245 cfm at 0.500” valve lift, but the air flow continues to increase as the valve opens further … 275 cfm at 0.600” valve lift, and 290 cfm at 0.700” valve lift. A motor equipped with unported 4V cylinder heads and a “street cam” opening the intake valves 0.600” off their seats has the potential of producing 550 horsepower! After porting the air flow performance further improves ... 290 cfm at 0.500” valve lift, 325 cfm at 0.600” valve lift, and 350 cfm at 0.700” valve lift. That much air flow is equivalent to approximately 700 horsepower.
There are two practical differences between motors equipped with 2V heads and motors equipped with 4V heads one must consider in deciding which type of cylinder head to use; (1) differences in power characteristic between the two motors and (2) differences in gearing required by the two motors. The average cross-sectional area of the 2V intake port is a bit smaller than the average cross-sectional area of the 4V intake port, thus the 2V intake port is tuned for a power band about 1000 rpm lower than the power band of a motor equipped with 4V cylinder heads. Some street performance and sports car enthusiasts prefer the low rpm biased power characteristic of the smaller cross-section 2V ports. The mid-range rush of an engine employing 2V cylinder heads is not as pronounced as an engine with 4V heads therefore the power characteristic of a 2V motor seems somewhat more sedate. Those who prefer the power characteristic of a motor with 2V heads feel it is more "refined" while the 4V motor is "brutish". A motor with 2V heads also lacks the endless high rpm pull of a motor with 4V heads. The 2V cylinder heads are better choices for low rpm torque applications such as towing and hauling vehicles, off-road and rock crawling vehicles and heavy vehicles in general.
The Cleveland 4V cylinder head design was chosen by Ford because of its suitability for endurance racing, it provides an unprecedented wide and flat torque curve and a power band characterized by a strong mid-range rush. Power at lower rpm with the 4V cylinder head is peppy, smooth and linear. As long as it is not "over-cammed" the low rpm power of a 4V motor is every bit as strong as the low rpm power of a 2V motor. A 4V motor can "light-up" the rear tires at low rpm with little effort. A 4V motor builds "steam" with engine speed and then hits a strong mid-range rush which is like the afterburners of a jet engine kicking-in! From that point on a 4V motor pulls harder and harder as the engine speed climbs (as long as the carburetor is big enough). I call the 4V motor's power characteristic the "Charging Rhino". Although I personally do not have a desire for a motor producing more horsepower than what can be achieved with the 2V cylinder heads, I prefer the power characteristic of a motor equipped with 4V cylinder heads.
Due to this difference in power characteristic 4V motors require lower gearing to operate the motor better within its power band which is at generally higher engine speeds. Since 2V motors do not require as low gearing as 4V motors they will potentially cruise at a more comfortable (lower) engine speed and return better fuel economy. Newer 5 and 6 speed manual transmissions and 4 speed automatic transmissions can provide both low gearing for acceleration and high gearing for cruising and fuel economy; the 4V motor is at less of a disadvantage in regards to "cruise" engine speed and fuel economy when used in conjunction with 4 speed automatic transmissions and 5 or 6 speed manual transmissions. Conversely, in that regard the 2V motor is at more of an advantage with 2 or 3 speed automatic transmissions and 3 or 4 speed manual transmissions.
In summary, as they came from the factory both versions of the 351C had good low rpm power, good throttle response, and good drivability. The 351C 4V power band/torque curve was wider, with performance to about 6000 rpm. The volumetric efficiency of the 2V cylinder head is excellent, and the 4V cylinder head has tremendous capabilities in that regard. With ratings of 250 horsepower (351C 2V) and 300 horsepower (351C 4V) the volumetric efficiency of both versions of the 351C was under-utilized. Higher engine speeds and higher outputs are attainable with relatively minor modification of the engines’ induction and exhaust systems.
A PRIME EXAMPLE
There is a forum member whose Pantera motor makes over 400 bhp at 6000 rpm on a chassis dyno; i.e. over 500 bhp at the crank. The engine is equipped with the standard stroke factory crankshaft. The iron 4V heads are ported, the compression ratio is set at 10:1, the induction is individual runner with fuel injection, the exhaust is a bundle of snakes type exhaust. The camshaft is a relatively mild hydraulic roller cam, the Crane HR216 camshaft. The cam specs are:
278°/286° advertised duration
216°/224° duration at 0.050"
0.562"/0.586" valve lift
112° lobe separation angle
The camshaft's specs are rather modest by today's standards. This is actually a very mildly tuned engine with good drivability, yet the power output is tremendous. Since the cylinder heads were designed to make peak horsepower at 6000 rpm the engine didn't need a long duration camshaft to make high rpm power. A lot of money was invested in the induction and exhaust systems, and that investment paid-off. The volumetric efficiency of the iron 4V heads is very high with IR induction, and the bundle of snakes exhaust helps too. This exemplifies the advantages of a good induction and exhaust system and it exemplifies the capabilities of the 4V cylinder heads. It also exemplifies the concept of improving the engine's performance without diminishing the power characteristic in any aspect. The "state of tune" was increased by improving the induction system, not by using a "big" camshaft. The design of the 4V cylinder heads gives us this option.
This naturally aspirated 500+ horsepower 351C (5.75 liter) makes as much horsepower as Chevrolet's naturally aspirated 7 liter LS7 Corvette engine, which is touted by Chevrolet as being a show case of high technology. The LS7 is 35 years newer, it uses 11:1 compression, CNC ported heads, a higher lift hydraulic roller camshaft and induction via a modern long runner fuel injection intake manifold.
The output of this 500+ horsepower naturally aspirated 351C is less than 50 horsepower lower than the output of the engine in the Ford GT. Like the LS7, the engine in the Ford GT is 35 years newer than the 351C. It displaces 5.4 liters, it is equipped with 32 valve dual overhead cam cylinder heads, high lift camshafts, and induction via a supercharger putting out 12 psi of boost!
Unlike either of these engines, the 351C was mass produced as inexpensively as Ford was capable, and it was manufactured from lowly cast iron. Yet if you bolt on a modern high-lift low-overlap camshaft (112° to 114° lobe centers, 50° to 60° overlap) and good induction and exhaust systems it can still keep up with these highly touted engines, and it can do so while maintaining its wonderful power characteristic. I consider 500 horsepower a high state of tune for a 351C, but it doesn't need to run like a drag-race engine to achieve high output ... unless of course you want it to.
PART 4 - THE COMBUSTION PROCESS (thermal efficiency)
The aspect which most significantly contributes toward the power producing capabilities of any engine is the combustion process. The combustion process is therefore the logical place to begin when optimizing an engine or tuning it for higher output. The parts which influence the combustion process include the design of the combustion chamber, the design of the piston dome, the compression ratio, the seal of the piston rings against the cylinder wall, the seal of the valves on the valve seats, the cylinder to cylinder consistency of the fuel air mixture, the quality of the fuel air mixture (better fuel atomization), the cylinder to cylinder consistency of the ignition system and the quality of the spark produced by the ignition system.
The shallow poly-angle combustion chamber of the Cleveland cylinder head is a very good design. It is the most important aspect of the head and the single biggest contributor towards the power producing capabilities of the Cleveland engine, yet people seldom think twice about it. The biggest improvement aftermarket cylinder heads offer over the factory cylinder heads is their “high swirl” combustion chambers.
COMPRESSION RATIO IMPROVEMENT
If the 351C in your car is a low compression version (having open combustion chamber cylinder heads and less than a TRUE 9.5:1 "static" compression ratio) then the single most important step to take towards improving its performance shall be to raise the engine’s compression ratio. Raising the compression ratio of any low compression 351C to operate on 91 octane (US/Canadian) pump gasoline (nominal 10:1 static compression ratio) shall increase the engine’s output by about 20 horsepower and improve the engine’s responsiveness. Although this does not sound like a lot of horsepower the amount of snap added to the engine’s performance gives the impression that the horsepower has increased much more than it actually has. It is not possible to set-up an engine having low compression to provide the type of performance most people are looking for. You are not giving the standard displacement 351C equipped with 4V cylinder heads a fair opportunity to show you the performance it is capable of having if you don’t raise the compression ratio.
The compression ratio of an engine is limited by the octane of the gasoline that shall be used. There are two rating systems for octane being used around the world, many people are unaware of the differences and unaware that low octane "regular" fuel in Europe is high octane "premium" fuel in the US and Canada. Since this is an international forum, its important that we are all on the same page, the information below should help.
Although we usually refer to compression ratio in terms of the “static” specification, it’s the “dynamic” compression ratio that more accurately describes an engine’s operating compression ratio, and therefore more accurately describes the limitation in the amount of compression a motor can tolerate. The “dynamic” compression ratio takes into account the piston’s “dwell” time at bottom dead center and how many degrees after bottom dead center (ABDC) the intake valve closes. My preference is to set-up an engine to operate on 91 octane US/Canadian pump gasoline (equivalent to gasoline rated 95 octane everywhere else in the world) since higher grade 93 or 94 octane fuel is not universally available throughout North America. The factory “Cleveland” cylinder heads, whether they have quench style combustion chambers or open style combustion chambers, can tolerate a maximum of roughly 8:1 dynamic compression with 91 octane US/Canadian pump gasoline. The aftermarket cylinder heads which are cast in aluminum and equipped with high-swirl combustion chambers are capable of tolerating at least 8.4:1 dynamic compression ratio burning the same pump gas. My preference is to set the dynamic compression of a street engine a little lower than the maximum amount in order to give the engine a margin of safety. 10:1 static compression combined with a camshaft which closes the intake valve at 70° ABDC results in a dynamic compression ratio of 7.66:1. That’s a reasonable margin of safety for the factory (iron) heads. For your comparison, the factory 351 Cleveland engines with the highest static compression ratios, i.e. the 1971 BOSS 351 and the 1970 351 4V, had dynamic compression ratios of 7.69:1 and 7.62:1 respectively.
Following are seven common scenarios for raising the compression ratio of the 351C.
A Word of Caution: Decking the block, milling the heads, installing pistons with greater compression height, or installing pup-up dome pistons increases the importance of checking piston to valve clearance during assembly of the motor, especially with high lift or long duration camshafts. Pop-up dome pistons will normally require more total ignition advance too.
Piston rings are a subject of rapid technological development fueled by the auto manufacturer’s pursuit of better fuel economy and lower emissions. It’s an area that you should spend some time researching and getting advice before you make a purchase. Piston ring thickness and tension creates friction that resists crankshaft rotation. Decreasing piston ring thickness or tension reduces the energy required to keep the crankshaft rotating, and therefore increases the power available at the rear wheels. The additional cost of high-tech rings over the price of standard cast iron rings constitutes one of the least expensive ways to increase horsepower. Modern thinner or lower tension piston ring sets can offer higher output at the rear wheels and better durability with no penalty in the life of the rings. A good quality OEM thickness 5/64” plasma moly ring set using a barrel faced ductile iron top ring will cost about $100 to $120. A top-of-the-line "thin" 1/16” chromium nitride faced ring set using a steel top ring will cost $280 to $380. So improving the technology of the piston rings will cost $280 or less. That's some relatively inexpensive horsepower. The Ross pistons I recommend are designed for the thinner 1/16" rings.
IGNITION SYSTEM IMPROVEMENT
If the 351C in your car is equipped with a breaker point ignition, or even an old and tired breakerless ignition then it must be improved. A smooth and precise operating "high output" breakerless ignition system is just as essential to a high performance engine as raising the compression ratio. There are many ignition systems to choose from. One possible ignition system choice that employs Ford parts is a Ford Duraspark distributor calibrated for 20° of centrifugal advance in by 3000 rpm, triggering a Duraspark I module and a Duraspark I coil.
California was the only state in which Ford vehicles were equipped with Duraspark I ignitions; Ford enthusiasts outside of California were not aware of the existence of this ignition and had no experience with its performance, which was a shame. The Duraspark I ignition was also known within Ford as the "high output ignition", it was much more than a different Duraspark module, it was considered an entirely different ignition system than the Duraspark II ignition system. It was Ford's first high output ignition system, and Ford's first ignition system to employ "dynamic dwell". The Duraspark I ignition was utilized in all California V8 equipped applications in 1977, and limited to California 302 V8 applications in 1978 and 1979. The Duraspark I ignition module is identified by its RED wiring sealing block.
At the heart of the Duraspark I ignition was a special ignition coil having a very low primary winding resistance. The coil was also operated with no ballast resistance; therefore current flow in the primary windings was substantially increased in comparison to the primary current of Ford's standard (Duraspark II) ignition system coil. The core of the Duraspark I coil was designed to accept a much higher magnetic charge from the increased current flowing in the primary windings, thus producing a substantially higher voltage to the spark plugs. The higher magnetic charge also allowed the coil to reach "full charge" more rapidly than Ford's Duraspark II ignition system coil. Spark intensity was greatly increased ... especially at higher rpm. If this coil's primary winding had been charged with the conventional "fixed-dwell" control utilized by the Duraspark II electronic ignition system it would have overcharged at low rpm and overheated. Therefore an ignition module with a unique primary current control circuit was required to compliment this coil.
Differing from the various Duraspark II ignition modules, the Duraspark I module didn't control charging of the coil in the conventional way. The Duraspark I module utilized dynamic dwell, meaning the module constantly adjusted dwell based on current flow in the coil's primary circuit, independent of engine speed. This prevented over charging or under charging the coil throughout the motor’s rpm range. Dwell therefore varied with respect to the degrees of crankshaft rotation but remained relatively constant with respect to actual coil charging time; and the coil was properly charged throughout the engine's operating range.
The Duraspark I ignition produced the most consistent and most potent spark of any Ford ignition. This was Ford’s best ignition for igniting lean mixtures or rich mixtures, which was the purpose for its existence. The ignition ignited mixtures the Duraspark II ignitions could not. The dynamic dwell feature gave this module good high rpm performance too as the coil was charged properly (never under-charged or over-charged) from idle to 7000 rpm. This ignition’s design was more elaborate than the design of the Duraspark II ignition, and therefore it was more costly for Ford to manufacture (replacement Duraspark I modules cost several times the price of replacement Duraspark II modules).
continued in the next post
PART 5 - INDUCTION SYSTEM (volumetric efficiency)
Feeding air and fuel to an 8 cylinder engine via a single carburetor is efficient in terms of cost and simplicity but a bad design in terms of high performance. Cylinder to cylinder consistency of the fuel air mixture is difficult to achieve with a single carburetor, and it is greatly influenced by intake manifold design. The manifold design must insure that equal amounts of air and fuel are flowing into each cylinder, it must not do anything that may cause the fuel to fall out of suspension, and it should prevent or control fuel puddling. The quality of the fuel air mixture is greatly influenced by carburetor design. Improved atomization of fuel (smaller fuel particles) insures the fuel stays in suspension, distributes more equally cylinder to cylinder, and ignites more readily in the combustion chamber. This is one reason why you shall find I emphasize the use of carburetors with annular booster venturis.
An individual runner induction system alleviates these concerns. An individual port fuel injection system employing a manifold with long equal-length runners also alleviates these concerns. A throttle body fuel injection system mounted on an intake manifold originally designed for a carburetor atomizes the fuel excellently, but it does not alleviate the concern for cylinder to cylinder consistency, fuel puddling, etc. I shall touch on these alternatives in this section.
Things we can do to improve the performance and/or volumetric efficiency of the induction system include increasing the size of the carburetor, improving the calibration of the carburetor, selecting a carburetor which atomizes fuel better, improving air/fuel mixture flow through the intake manifold, smoothing the intake ports, grinding 3 angle intake valve seats, improving the camshaft lobe profile, and increasing valve lift.
In practical terms those things boil down to a larger carburetor with annular booster venturis, an aluminum intake manifold (Ford, Edelbrock, Blue Thunder) or reworking the factory cast iron intake manifold to accommodate the larger carburetor, and a camshaft that lifts the valves further open with matching valve train improvements. Key valve train improvements include single groove stainless steel valves, higher rate valve springs, stiffer-thick wall push rods, and improvements to the factory rocker arms. The subject of valve train improvement shall be discussed in additional detail in the valve train section which follows this section.
SINGLE FOUR BARREL CARBURETOR INDUCTION
Four Barrel Carburetors
At 6000 rpm a 351 cubic inch motor would theoretically inhale 609 cubic feet of air per minute if the volumetric efficiency were 100%. At 7000 rpm the same motor would inhale 710 cubic feet of air per minute assuming 100% volumetric efficiency. Assuming 90% volumetric efficiency a 351 cubic inch motor will inhale air at the rate of 548 cfm at 6000 rpm or at the rate of 639 cfm at 7000 rpm.
However, as the volumetric efficiency of a motor improves the intake manifold vacuum at wide open throttle shall decrease. The intake manifold pressure of a motor with 100% volumetric efficiency is theoretically equal to atmospheric pressure at wide open throttle. The airflow rating of carburetors is measured at a fixed depression, such as 1.5 inches of mercury in the case of Holley carburetors. If the depression across a Holley carburetor is less than 1.5 inches of mercury at wide open throttle it will not flow the amount of air it is rated at, the motor shall require a carburetor with a larger rating than what we calculated in order to supply adequate airflow at 6000 or 7000 rpm. The reason for this is not because the motor demands more air flow than what we calculated but because the carburetor, which is rated at a depression of 1.5 inches of mercury, flows less air if the depression is less than 1.5 inches of mercury; in other words the flow rating of a carburetor as determined at 1.5 inches of mercury becomes less relevant as the volumetric efficiency of a motor increases. For any given quantity of air flowing into the engine a larger carburetor will require less intake manifold vacuum to supply that quantity of air, therefore the intake manifold vacuum at any given rpm shall be less and this allows for higher volumetric efficiency.
Both the 351C 2V and the 351C 4V have higher volumetric efficiency than the popular in-line-valve V8s people are more familiar with; at wide open throttle the vacuum in their intake manifolds will drop lower than it does in those other V8s if the carburetor is large enough to allow it. This is the reason larger carburetors are recommended for the Cleveland engine series. If an owner selects parts for the 351C induction system following the same guidelines people follow when selecting parts for a SBC or SBF the 351C shall not perform any stronger than a SBC or SBF; the superior volumetric efficiency for which the 351C is known shall be quenched. Contrary to the carburetor sizing conventions you may be familiar with the 351C (especially the 4V version) is designed to inhale more air than other engines and it responds well to a bigger carburetor. There is no penalty in drivability or throttle response as long as the carburetor is calibrated properly.
On top of that the 351C 4V is capable of operating over an extraordinarily wide power band, certainly wider than any other OHV engine from its era. The first 351C 4V performance manifolds designed by Ford were designed for list #4575 Holley Dominator carburetors (1050 cfm)! Ford’s earliest carburetor recommendations also included the Holley 850 cfm double pumper. The 351 Cleveland engines require carburetors designed for engines having higher volumetric efficiency and in the case of the 351C 4V a wide power band too. The usual carburetor choices for a 351C 2V usually range from 650cfm to 750cfm; for the 351C 4V those choices usually range from 750cfm to 850cfm. None of these carburetors are too big for a 351C street motor, especially if they are equipped with annular booster venturis. With a 351C 4V street motor it is a challenge to find a carburetor that performs well at low rpm while also being large enough to take advantage of the WOT (wide open throttle) volumetric efficiency of that motor.
Annular booster venturis atomize fuel better and provide a stronger fuel metering signal at low air velocity. In other words, annular booster venturis benefit the low rpm and mid-rpm performance of a motor in the same manner as the smaller primary throttle bores of a spread bore carburetor. These attributes make annular booster venturis popular for improving the low rpm operation of performance engines, where they have earned a reputation for improving torque, horsepower and throttle response at low engine speeds. However the improvement in fuel atomization distributes fuel more consistently throughout an intake manifold, resulting in more consistent fuel/air ratio from cylinder to cylinder, therefore annular booster venturis actually improve torque and horsepower across a motor's entire power band; and they improve fuel economy too! The only drawbacks of annular booster venturis include their larger physical size (which reduces the airflow capability of a carburetor by a relatively small amount) and their greater cost of manufacture.
If an owner selects a smaller carburetor it’s not the end of the world. A 600 cfm to 650 cfm carburetor is fine for daily transportation purposes and even a playful bit of acceleration from time to time. But I don't recommend that choice for the performance minded owner. A 351C equipped with a smaller carburetor will flatten out sooner when accelerating and lose the eagerness to rev far beyond 6000 rpm. The smaller carburetor may lower the rpm at which peak horsepower occurs and it shall definitely impair the engine’s volumetric efficiency.
The most important aspect of any carburetor is not its size but how well it has been calibrated to suit your car’s motor. Regardless of what size carburetor you choose, if it is calibrated poorly the motor shall perform poorly, if it is calibrated well it shall perform well. An engine equipped with a well calibrated 600 cfm carburetor will make more horsepower than if it were equipped with a poorly calibrated 750 cfm carburetor; but if the 750 cfm carburetor is calibrated as well as the 600 cfm carburetor then the motor will perform better with the larger carburetor. It is rare to find an out-of-the-box carburetor that is calibrated 100% ideally for your application. For this reason many enthusiast prefer a carburetor having features that make it easier to tune.
Single Plane 351C Intake Manifolds
Dual Plane 351C Intake Manifolds
A dual plane intake manifold is the best choice for a street driven vehicle in terms of overall functionality and usable performance. Most dual plane intake manifolds will blunt an engine’s power in the upper rpm region, but they improve a motor’s “grunt” at low rpm, they improve drivability and they perform best in the rpm range encountered 90% of the time when driving on public roads. Dual plane intake manifolds usually improve vacuum at idle by at least 50%; achieving the best possible manifold vacuum allows the positive crankcase ventilation system (PCV), the distributor vacuum advance mechanism, and the power brake vacuum booster to all work within their intended design limits for best possible operation (automatic transmission vacuum modulators too).
The 4V intake port entrance is about 2-1/2" tall, yet the runners in the factory intake manifolds are not nearly that tall, they are closer to 2" tall. The factory intake manifold runners flare open to match the height of the 4V intake port entrance. The gas flow in the factory induction system starts in a runner with a cross sectional area of about 3 square inches, then it expands to a cross sectional area over 4 square inches at the intake port entrance, then past the intake port entrance it returns to a cross-sectional area closer to 3" again. This is not ideal. One way to achieve an induction system having a more consistent cross-sectional area is to use the Blue Thunder manifold which has full height runners that complement the opening of the 4V intake port; the Blue Thunder manifold was designed to be a "wide open induction system" manifold thus complimenting the engineering of the 4V intake port, it performs very well with the iron 4V heads.
Like the factory intake manifolds, the runners of the Edelbrock Performer manifold and Scott Cook’s dual plane manifold are about 2” tall; however the runners of the Edelbrock and Scott Cook manifolds do not flare open, they are designed to achieve a more consistent cross-sectional area; therefore the runners are smaller than the 4V intake port entrance. Those manifolds can be mated to the 4V heads as-is. Do not blend the runners of the manifolds to match the opening of the 4V intake port; they perform better if they are not blended. If you wish to eliminate the mismatch and make the cross sectional area more consistent the proper way to do so would be to fill the inlet of the 4V intake port to match the Edelbrock or Scott Cook manifold runners (this is also called stuffing the intake port). Filling the inlet of the intake port about 1/8" on the left side, and about 1/2" on the floor, gives the intake port a more consistent cross-sectional area (the average cross-sectional area is reduced from 2.9 square inches to about 2.7 square inches) and makes the port smoother, eliminating the push-rod bump and ramped floor built into the port’s entrance. Of course, Scott Cook’s cylinder heads feature stuffed 4V intake ports out of the box.
One warning however, the 4V intake port was intentionally flared open, creating the ramped floor and push-rod bump at the port’s entrance, in order to incorporate features which increase air flow within the port. In spite of the fact they create an irregularly shaped port the features work very well at increasing flow. Intake manifolds which lift the gas flow within the port so as to avoid these features may actually result in a decrease in gas flow or engine performance. Stuffing the port will definitely decrease gas flow. If you're going to stuff the port entrance then the port should be "ported" further within afterwards to regain the lost port volume, to make the ports cross-section and shape smoother and more consistent, and to regain the flow that was lost by stuffing it. The port actually works very well "as-is", it doesn't require "fixing". This is why I recommend the Blue Thunder manifold which allows the intake port to operate optimally in the way it was originally designed to do so.
Fuel Supply System
It’s very likely the fuel supply system of your car’s engine will also require reworking in order to supply sufficient fuel for the motor’s higher output. The Robb Mc Performance #1020 mechanical fuel pump is rated for up to 550 horsepower. Plumb the fuel system in metal tubing as much as possible, keep the hose sections as short as possible, use a tubing bender to put smooth large radius bends in the metal tubing, avoid 90° tubing fittings. Plumb the pump suction in ½” (AN-8) tubing or hose and plumb the pump discharge in 3/8” (AN-6) or ½” (AN-8) tubing or hose. Install a high flow fuel "pre-filter" designed for the fuel pump inlet (75 to 150 micron) and install a high flow fuel "post-filter" designed for the fuel pump outlet (10 microns for fuel injection or 40 microns for a carburetor). RobbMc Performance
Pantera owners: the tubing in the fuel tank that supplies the fuel pump is only 5/16" OD; this is much too small. If you're building a high output motor for a Pantera, this is one item that shall require modification. Upgrade it to 1/2" tubing.
The picture below details the proper way to plumb a fuel system using an electric fuel pump for both carburetors or fuel injection.
INDEPENDENT RUNNER INDUCTION
Another induction system modification worth mentioning is an independent runner induction (often abbreviated IR) composed of a Weber 48IDF intake manifold manufactured by Aussie Speed of Australia and four Weber 48IDF two barrel down draft carburetors. The Aussie Speed manifold is designed as a 2V manifold, but it is also designed to seal-up the larger intake port openings of a 4V cylinder head.
The Weber IDF carburetor is Weber’s most popular two barrel down draft carburetor for racing, high performance automobiles and sports cars. It has been used as OEM equipment in a few limited production vehicles, including a Ford Escort (the European Ford Escort RS2000 Group 1 cars). It is accepted by very many sports car hobbyists as a suitable replacement for various Delorto and Solex carburetors. Like the side draft Weber DCOE carburetor, the Weber IDF carburetor is sold by Pegasus Racing, which is an indicator of the carburetor’s popularity. There is no demand for Weber’s other two barrel down draft carburetor, the IDA carburetor, outside of the American muscle car niche market. Weber's IDA carburetor is not as well suited for street applications as their IDF carburetor.
The IDF carburetor is offered in 40, 44 and 48 mm bore sizes. The main, idle, air correction and accelerator pump jets, the emulsion tubes and venturis, are interchangeable. It has a float design that makes it very popular for off-road applications, a vacuum advance port, and four progression holes for smooth light-accelerator response. The differences between the IDA and the IDF, like two additional transfer circuits, add up to make a big difference in the IDF’s part-throttle performance and its suitability as a carburetor for a year-round daily-driver application.
An IR induction such as this is more expensive to purchase, it is more time consuming to tune, and it often requires more frequent maintenance. However, in comparison to a single four barrel carburetor induction system the benefits of an IR system include quicker throttle response, faster acceleration, a wider power band and substantially improved volumetric efficiency. Aussie Speed
Fuel injectors atomize fuel better than a carburetor at low engine speeds and normally account for a 10% to 15% improvement in torque at low rpm. Port fuel injection also eliminates the issues of fuel falling out of suspension, fuel puddling, and uneven fuel distribution associated with intake manifolds designed for carburetors.
Throttle body fuel injection shall always be a viable option for retrofitting fuel injection to an older engine in terms of simplicity, cost and stealth because it can be installed in place of an existing carburetor, and therefore it does not require the replacement or modification of an intake manifold, it takes up no more room in the engine compartment, and it even uses the same air filter assembly.
However the big news in fuel injection for the 351C is the two port-fuel injection intake manifold kits manufactured by Trick Flow® Specialties. Both of these intake manifold kits are an overwhelmingly better way to fuel inject a 351C in terms of performance compared to utilizing an intake manifold designed for a carburetor.
The R-Series intake is a long equal length runner design which is tuned for high performance street engines. Combining long equal length runners with the improved fuel atomization at low rpm typical of fuel injectors this manifold has the potential to boost the lower rpm power of a 351C much in the same way this type of manifold boosted the lower rpm power of the 5.0 HO V8 in the 1980s. This manifold provides a superior way to achieve the type of low rpm grunt that people are trying to duplicate when they build stroker engines. The R-Series intake manifold is claimed to have 13.3” long runners and an overall height of 11.000 inches. It is available with either a 75mm throttle body inlet or a 90mm throttle body inlet.
The Box-R-Series intake features a large plenum/short equal length runner design which maximizes mid-to-high-rpm power making it ideal for racing applications. This manifold has the potential to improve volumetric efficiency like an individual runner induction system but only requires one throttle body. The Box-R-Series intake manifold has a 90mm throttle body inlet, and an overall height of 12.000 inches.
Both Trick Flow® EFI manifold uppers are offered in a choice of silver and black powder coated finishes or natural aluminum which allows the customer to finish the manifold as they prefer. The common base used for these manifold kits is finished in bare aluminum, the port outlets are 2.100 inch x 1.500 inch at the cylinder head; and it is designed to bolt up to both 2V and 4V cylinder heads. Trick Flow Specialties
CYLINDER HEAD PORTING AND PORT STUFFING
You may consider having some minor work performed on the heads ... pocket clean-up and 3 angle valve seats, plus a little work in the roof and sides of the exhaust port. Its hard to give guidance regarding selection of a business to port 4V cylinder heads in broad terms. Do not agree to extreme porting of the 4V heads unless the business has a decades old reputation for porting 351C 4V cylinder heads; such as Koontz and Company (Arkadelphia Arkansas) or Valley Head Service (Northridge California). Most "cylinder head porting businesses" do not understand the 4V heads. I've seen 351C 4V performance worsened by many businesses claiming to be professionals at cylinder head porting. It is healthy to be hesitant and cautious about handing over your 4V cylinder heads to any business for modification. A simple amount of pocket and port clean-up combined with 3 angle valve seats will increase air flow through both the intake and exhaust ports by 50 cfm at 0.600" valve lift.
More to Come - Under Development
PART 6 - EXHAUST SYSTEM
Things we can do to improve the exhaust system include increasing valve lift, improving the camshaft lobe profile, grinding 3 angle exhaust valve seats, smoothing and moderate porting of the exhaust ports, installing steel tubing headers and an unrestrictive (low back pressure) exhaust system. A low back pressure exhaust system includes 2-1/4” to 2-1/2” intermediate pipes (with an H-type or X-type cross-over if possible), free flowing mufflers, and tail pipes that are as short as possible (dumping ahead of the rear tires in front engine cars). An exhaust system can only perform as well as its weakest link. As mufflers become less restrictive the gains that can be realized by improving other aspects of the exhaust system multiplies. The higher the output of an engine the higher the gains that may potentially result from exhaust system improvements.
STEEL TUBING HEADERS
The first rule of thumb: the inside diameter of the primary tubes should be approximately 110% the diameter of the exhaust valve. That rule of thumb equates to headers with 1-7/8” to 2” primaries for iron 4V heads or 1-3/4” primaries for all other heads. Another rule of thumb: low rpm power is usually improved by smaller and longer tubes & high rpm power is usually improved by bigger and shorter tubes. The design of the “collector” is also very important.
Standard “4 into 1” steel tubing headers should feature primary tubes of reasonably equal length, they should be made of tubing of the proper diameter and functional length, they should be shaped with no sharp bends, and they should terminate into a properly designed collector. 180 degree exhaust systems (aka bundle of snakes exhaust systems) combine primary tubes from both banks of cylinders based on the engines firing order, in order to achieve the maximum possible separation between exhaust pulses in each collector. The goal is to broaden the tuning of the exhaust system thus improving mid-range power over a wide range of engine speed. Tri-Y headers (aka 4 into 2 into 1 headers) pair the cylinders on each bank of the engine so as to provide the maximum separation between exhaust pulses per bank. It is a more practical design for achieving performance goals similar to a bundle of snakes type exhaust system without taking up as much engine compartment space. With a Tri-Y header I've been told the SECONDARY tubes are a more important design aspect than the primary tubes. In other words when space is limited it is better to shorten the primary tubes and lengthen the secondary tubes.
After those considerations the most important aspect of header design (and the exhaust system as a whole) is the tubing geometry; i.e. construction details such as bend radii, intersection angles, nozzle angles, and diffuser angles. These details have a tremendous impact on the performance of the exhaust system. Any header (and exhaust system) has to make concessions in its construction to clear various parts of the chassis, the suspension, the steering, the starter motor, the bell housing or the transmission. Each one of those concessions also has the potential to inhibit the theoretical pulse and reflection behavior of an exhaust system and negatively impact the exhaust system's performance; thus a “real world" exhaust system does not always perform as well as expected. One particular aspect of header geometry was discovered long ago to greatly impact the performance of the 351C 4V exhaust port ... the 4V exhaust port works better if the primary tubes extend straight out of the head as far as possible.
The Pantera chassis does not provide enough space for headers with primary tubes of the proper length. Nor does it provide enough space for decent mufflers. It is possible to install a cross-over under the car connecting the left and right sides of the exhaust system, but it is almost as long as the intermediate pipes! The one good thing about the Pantera chassis, it allows enough space on either side of the engine so the primary tubes can extend straight out of the head for several inches.
The European GTS exhaust system should be considered a minimum upgrade for Panteras equipped with the "small tube" factory exhaust system. The GTS system is reasonably quiet, with a nice low frequency burble. The price of the system is reasonable, it’s easy to install and it has the all important factory look. The headers have the proper size primaries for iron 4V heads and the header flanges are nice and thick. They are designed as a "pseudo-Tri-Y" header lacking secondary tubes! The single collector is also too small. But surprisingly the headers perform better than they have any right to do so. The intermediate pipes are made of thick wall 2-3/8" OD steel tubing, they weave their way through the Panteras suspension perfectly. The system's biggest drawbacks are the mufflers which impact horsepower output above 5500 rpm. To work best with the Ansa mufflers your Pantera's motor should employ a camshaft with limited overlap (50° to 62°) which opens the exhaust valve early (80° BBDC or earlier).
The spaces provided for mufflers in the narrow body Panteras are only 10” x 10”, but the wide body Panteras have room for mufflers up to 16" in length because the rear tires are spaced outwards. I am not aware of a quiet, free breathing replacement for the restrictive, 9-1/2" x 9-1/2" Ansa mufflers, but there are two loud universal aftermarket mufflers that will fit in the same small spaces:
PART 7 - CAMSHAFT AND VALVE TRAIN
The large and heavy 351C 4V intake valve, the high ratio (1.73:1) rocker arm, and the canted valve geometry which splays the pushrods apart at extreme angles, constitutes one of the toughest valve train applications of any OHV engine. 500 to 570 horsepower was pro-level horsepower for any size endurance racing engine when the 351C was designed. In those days valves lifted off their seats by 0.550” to 0.600" was state of the art for a pro-level endurance racing engine. Although that's standard valve lift for a modern street performance cam, in the early 1970s solid flat tappet racing cams having 0.550” to 0.600" valve lift were very long duration - high overlap camshafts which pushed the limits in race engine performance and reliability. The valve train performance of our modern street engines was achieved via advancements in successive generations of camshaft grinding machinery and it was made usable by advancements in valve spring technology.
If you dramatically improve the valve train performance of your 351C engine, you have to assume you cannot take any short-cuts in the quality of the valve train componentry you select. Keep in mind your engine's modern valve train may be lifting the valves off the seats as much as racing cams did 40 years ago, with less camshaft lobe duration, and with hydraulic tappets rather than solid tappets!
The performance of the 351C 4V can be described as having good drivability, a strong dose of mid-range power, and the willingness to rev to high rpm. As I stated previously I consider this characteristic ideal, it is something I try to avoid diminishing in any aspect when I tune the engine for higher output. The camshaft plays a major role in that.
The other small block in-line valve motors people are more familiar with have small ports and small valves for the given displacement of the motor, thus pushing the motor’s power band into the very lower end of the rpm range. Such motors rely upon long duration camshafts and high rpm intake manifold design to widen the power band and promote mid and upper rpm power. The 351C 4V is just the opposite, the intake ports and valves were sized to give the motor a power band that peaks at approximately 6000 rpm and pulls strong out to 6500 rpm and in some cases even as high as 7000 rpm. The 351C 4V relies upon camshaft design (moderate valve event timing) and intake manifold design (dual plane) for its lower rpm performance. The wide and flat 351C 4V power band/torque curve was quite good for street performance off the show room floor, the power band simply does not need to be altered or raised.
351C 4V FACTORY CAMSHAFTS
If you're interested in one of the Ford camshafts, three of them are still available via the aftermarket.
HIGH LIFT RATE CAMSHAFTS
In selecting a camshaft to improve the induction system the goal is to find a camshaft which lifts the valves higher via higher-lift-rate lobes without straying too far from the factory camshaft timing. Straying too far from that timing will erode the engine’s good low rpm performance and drivability. The short intake duration of the factory camshafts proves a 351C with 4V heads does not need a lot of intake duration to have a high revving power band.
There are four common camshaft design errors that are made in regards to valve event timing that make things worse in regards to 351C 4V street performance:
Camshaft grinders are unwilling or incapable of grinding camshafts with LSA greater than 115°. In fact, camshaft grinders are even reluctant to grind cams with 115° LSA. The most “street-able” off-the-shelf aftermarket cams are ground with 112° to 114° LSA. Even when having a camshaft custom ground you’ll encounter less resistance from camshaft grinders if you specify 114° LSA instead of 115° LSA. For this reason, and this reason only, 114° is realistically the widest LSA we have to work with. Combining the capabilities of the 4V cylinder heads, camshaft timing within the limits I've recommended, and high-lift-rate camshaft lobes (max valve lift in the range of 0.550” to 0.600”, hydraulic intensity in the range of 50 to 62) imbues an engine with good drivability, a wide power band AND hard-hitting performance. This I guarantee.
One GOOD off the shelf cam sticks out in my mind, the Crane Cams HR-216 hydraulic roller cam.
If you would like my assistance in specifying a custom cam for your 351C street or sports car application you are welcome to contact me privately, via one of two methods:
(1) Via a "private message" using the messaging capability of these forums. i.e you'll have to join the forums.
(2) Via an email sent to "info at Pantera International dot org".
If you contact me via one of those two methods, I will gladly assist you.
I have assisted many people over the decades (since the 1970s), and I will gladly assist you as well. But frankly, I would prefer if you asked for guidance at the beginning of your project, rather than just asking for help with the camshaft.
Pay attention to this: I offer assistance to folks who own vehicles powered by the 351C who wish for nothing else beyond achieving the best performance the factory 351C castings had to offer.
I don't offer help choosing off the shelf cams, in terms of camshafts my specialty is penning custom cams for high performance street engines which perform better as street cams than any mass-produced cam available off-the-shelf. I don't offer help designing drag racing cams, I don't offer help designing cams for high output custom engines (beyond 450 bhp). Custom engines are those equipped with alloy heads and/or stroker crankshafts.
CAMSHAFT TIMING SETS
A high quality steel timing set having a 9 keyway crankshaft sprocket and a camshaft sprocket with steel teeth (as opposed to plastic teeth) is the durable and practical choice. The multi-index crankshaft sprocket is an invaluable aid in properly timing a camshaft. Some choices include: Roll Master #CS 3091, Ford Racing Performance Parts #M-6268-A351, or Cloyes #9-3621X9.
Unless your 351C has been specifically set-up to operate with low viscosity 0W or 5W synthetic motor oil, the recommended oil viscosity for Cleveland Fords (built to standard 351C specification) is 20W50, 15W40, 10W40 or 10W30.
Motor oil providing a high level of wear protection is required to prevent premature failure of flat tappet camshaft lobes, flat tappet lifter faces AND distributor drive gears. It is important to emphasize that installing a roller cam does not eliminate the need for motor oil providing a high level of wear protection; it is still needed for the distributor drive gear! The traditional recommendation has been to select oil containing more than 1200 ppm of both zinc and phosphorous, the constituents which make the anti-wear agent known as ZDDP. However a high level of ZDDP does not guarantee a motor oil provides a high level of wear protection. ZDDP oil additives do not help either; they reduce the wear protection properties of motor oil! My recommendation is 10W30 Valvoline VR1 Racing Oil, either petroleum based (silver bottle) or synthetic (black bottle); it is reasonably priced, it is readily available and it provides a high level of wear protection.
VALVE TRAIN PERFORMANCE, WEIGHT, AND LONGEVITY
The dynamic goal in a high performance valve train is to remain in control of the valves up to the motors rev limit. Parts should be rigid enough so that their shape does not distort. Parts should also be light, they should remain in contact with one another, and they should follow the cam's motion precisely. There should be no unwanted motion in the valve train; such as wiggling, bouncing, surging, floating or flexing. The properties of the moving valve train parts that work against the performance enthusiast are inertia, energy storage, flex, oscillation, resonance ... AND cheaply made parts!
Valve train wear increases proportionally to increases in valve spring force. Increasing valve spring force shall also lower the rpm at which hydraulic tappets collapse. We can’t increase a street motor’s valve spring force indefinitely if we expect the valve train to operate for many miles without needing rebuilding, or if we wish to avoid hydraulic tappet collapse. If we install the strongest valve springs recommended for street applications (such as those I’ve recommended below) and find valve float OR hydraulic tappet collapse occurring at a lower rpm than we prefer, or find the motor suffers from valve train instability issues, the next course of action may be to lighten the valve train.
Weight removed from a valve or valve spring retainer is more effective than weight removed from a push rod or tappet, due to the multiplication of movement built into the rocker arm. With a 351C, which has a 1.73:1 rocker arm ratio, any weight removed from a valve or valve spring retainer is 1.73 times more effective than the same amount of weight removed from a push rod or tappet. The most important characteristic for a push rod or tappet is to be completely rigid, free from flex and distortion. Since it is less effective to lighten these parts anyway, the prevalent reasoning is to choose these parts based on strength, and to focus on lightening the valve train via the valves and retainers, where each gram of weight reduction is more effective. A rule of thumb used in the hot rod industry says reducing the weight of these components by 1 gram will raise a motors rev limit by 25 rpm.
Manley’s stainless steel "severe duty" 4V intake valve weighs 139 grams. Manley’s stainless steel "severe duty" 4V exhaust valve weighs 108 grams. The intake valves each weigh 31 grams more than the exhaust valves! There is a lot of performance to be gained by replacement of the "severe duty" intake valves with lighter valves such as Manley's "Race Master" intake valves (129 grams). Reducing the weight of the intake valves by 10 grams raises the rev limit by 250 rpm. Adding titanium valve spring retainers for the intake valve springs is a moderately priced method for removing a few extra grams of weight, and it is complimentary to the use of light weight intake valves.
Accelerated valve seat wear and valve stem or valve guide galling are problems encountered by some racing engines employing titanium valves, however keep in mind that race engines employ very high lift rate camshaft lobes and very high valve spring forces. Race engine builders are also tempted to set the valve seats thin in order to improve air flow. Regardless of how many beat up titanium valves a race mechanic has in his tool box, the lower lift rate camshaft lobes, lower spring forces and wider valve seats utilized in high performance street motors are an ideal application for titanium valves, if they are needed and/or within the budget. Whereas Manley's Race Master valve will raise the rev limit by 250 rpm, a Manley titanium intake valve would raise the rev limit by 775 rpm!
There are not one but two reasons for replacing the OEM factory valves. (1) The factory valves have brittle heads; they sometimes crack near where the heads are induction welded to the stems. Cracking leads to the valve head falling off the valve stem while the motor is running, and destructive damage occurs to the motor. (2) The valve springs are retained by loose fitting multi-groove valve spring locks which are not fit for performance usage, i.e. higher than stock valve spring forces and high engine speeds. This is substantiated by Ford’s choice to install single groove style valves in the Boss 351. People have been replacing the factory Cleveland valves with Manley severe duty stainless steel valves for decades, since the motors were new. They are a high quality, time proven substitution. Manley Performance is located in Lakewood New Jersey; their telephone number is (732)905-3366.
Whatever brand of valves you choose, it is imperative the stainless steel or titanium valves you purchase have hardened steel tips. Cast iron or beryllium-copper valve seats are complimentary to stainless steel or titanium valves. To prevent rapid wear of stainless steel or titanium valves in the valve seat area the cylinder head’s intake seat width should not be less than 0.060” and the exhaust seat width should not be less than 0.080”; seat run-out should be 0.001” or less. Equip the cylinder heads with silicon-bronze valve guides to best compliment stainless steel or titanium valve stems. The valve stem to guide clearance should be set at 0.0010" to 0.0020" for the intake valves and set at 0.0015" to 0.0025" for the exhaust valves. Utilize spring loaded elastomer valve stem seals such as Ford Racing Performance Parts #M-6571-A50 or Manley Performance #24045-8; installation of this type of seal requires machining of the top of the valve guide to 0.530” diameter.
continued in the next post
Standard 351 Cleveland push rods are 5/16” diameter and 8.41" long, but when the block is decked, when the heads are milled, when factory head gaskets are replaced by gaskets having a different compressed thickness, or when parts like the camshaft, lifters, valves or rocker arms are changed the required length of the push rods shall change as well.
Push rod deflection can cause many seemingly unrelated engine performance problems; they are the weakest link in an overhead valve type valve train. It is important to use push rods in any application that are rigid enough for the spring forces, for the weight of the valve train components, and for the engine speeds involved. The canted valve Cleveland valve train splays the push rods off to either side of the intake port; these push rod angles expose the 351C push rods to angular bending forces not encountered in the valve train of in-line valve motors; the 351C needs a sturdier push rod. The push rod is not the appropriate component to use for reducing valve train weight or saving money. Using push rods that are “overkill” for their application is my way of insuring the push rods are perfectly rigid and there’s no possible way they contribute to any valve train related reliability or performance issues. Push rods should be manufactured from seamless chromoly tubing. The use of chromoly tubing alone will guarantee a more rigid push rod. The larger the OD of the push rod the more rigid it shall be also, increasing the wall thickness of the tubing does not increase push rod rigidity as much as increasing the outside diameter. Push rods being specified for hydraulic tappet applications should have a 0.040" restriction in one end to control the amount of oil flowing to the valve train. Of course, restricting oil to the valve train via the push rods is not a concern if a motor is equipped with tappet bore bushings having 0.060” orifices.
5/16” push rods made from 0.080” wall thickness tubing are considered adequate for a relatively stock motor but I recommend a more rigid push rod. 5/16” push rods with 0.105” wall thickness are a step up in rigidity. 5/16" push rods made from 0.116” to 0.120" wall thickness tubing are a favorite choice of mine for hydraulic flat tappet applications (spring force up to 330 pounds over the nose) because the passage in the middle of the push rod is only 0.072" diameter. The small passage acts as a restrictor to control the amount of oil flowing to the valve train in lieu of a restrictor in the push rod's tip. The most rigid recommendation however is a 3/8” push rod with 0.080” wall thickness; this has been a common recommendation for 351C applications for decades.
Smith Brothers of Redmond Oregon (800-367-1533) and Manton Pushrods of Lake Elsinore California (951-245-6565) are shops specializing in custom made push rods. Manley Performance Products and Trend Performance are also good places to shop for push rods.
The factory rocker arms are suitable for the hydraulic tappet applications being discussed. There are two common warnings in using the factory rocker arms: (1) Use only steel 4V fulcrums (the 2V fulcrums are made of aluminum). (2) Beware of factory rocker arms that have “lugs” along the edges immediately above either side of the fulcrum area. There is a problem with push-rod clearance when using those rocker arms with camshafts lifting the valves 0.550” or higher, therefore they should be replaced. Sealed Power #R-855 is a recommended replacement for the factory rocker arms.
Beyond those warnings the factory rocker arm has three potential weaknesses: (1) fulcrum bolt stretch, (2) push rod cup wear and (3) the quality of the valve stem contact patch (a rocker arm geometry issue).
Fastening the rocker arms to the pedestals with ARP #641-1500 bolts (4 packs) and #200-8587 washers (2 packs) is recommended to improve the strength of the fulcrum bolts and reduce the possibility of them stretching. The 1/8” thick washers are necessary because the ARP bolts are 1/8” longer than the factory bolts. With the fasteners thus improved the factory rocker arm is good for up to approximately 400 pounds over the nose and it can accommodate applications lifting the valves up to 0.615” off the seat. 0.615” valve lift was Ford’s recommended limit for the production rocker arms based on push rod clearance.
If you wish to upgrade to adjustable valve train and your motor’s factory iron cylinder heads are equipped with unmodified slotted rocker arm pedestals the Scorpion #3224 rocker arm can be bolted directly to the unmodified pedestal and provide push rod cup type valve lash adjustment. This is a high quality billet rocker arm that operates like an individual shaft mount rocker arm. Keep in mind the 5/16” fasteners limit this rocker arm to spring force of about 400 pounds over the nose.
If your motor’s cylinder heads are milled and tapped for 7/16” stud & guide plate type rocker arms the Yella Terra YT-6321 rocker arm is the hot tip. This very rugged rocker arm also performs like a shaft mounted rocker arm therefore it requires no studs, guide plates or hardened push rods. Internet pricing for the Yella Terra YT-6321 rocker arm is in the range of $785 US dollars for a set of 16.
The next step up in price is the T&D Machine individual shaft mount rocker arm, which is available in steel, this is its main benefit. Whereas billet aluminum rocker arms are good for about 10,000 miles, a steel rocker arm is a better choice for an engine planned for high mileage.
ROCKER ARM GEOMETRY
There are six variables which impact the geometry of a rocker arm; (1) the amount of camshaft lobe lift, (2) the design of the rocker arm, (3) the height of the rocker arm's fulcrum, (4) the rocker arm's lateral distance from the valve stem, (5) the height of the valve stem and (6) the length of the push rod. Optimum rocker arm geometry minimizes side thrust on the valve stem and guide which has two substantial benefits; (1) it minimizes the wear of parts AND (2) it minimizes the rocker arm’s contribution to oscillation induced valve train problems.
Geometrically ideal rocker arm geometry will set the rotational axis of the rocker arm at the same height (perpendicular) as the valve tip when the valve is 50% open. That’s just on the rocker tip side, there is also geometry on the push rod side, but getting close to the correct geometry on that side depends upon the rocker arm being designed with that as a consideration, and designed for the amount of lobe lift employed by your motor’s camshaft. When the geometry is correct on both ends the rocker arm will impart the most possible lift to the valve, this will not occur unless the geometry is correct at the rocker arm tip AND the push rod. This indicates the valve train is following the motion of the camshaft lobe most precisely, which is one of the primary goals of a high performance valve train.
Correct geometry at the rocker tip will place the sweep of the rocker tip nearest the rocker arm at fully closed and fully open, the sweep will be furthest from the rocker arm at 50% open, and the rocker tip shall be in the middle of its sweep at approximately 25% and 75% open. This geometry will always result in the narrowest sweep pattern, although there is nothing beneficial about a narrow sweep pattern, it is just a method of evaluating the rocker arm geometry. This description of sweep pattern will be in direct opposition to many of the rocker geometry instructions you shall run across. A few of the camshaft companies are notorious for promoting bogus rocker geometry instructions. The hot rod industry teaches home mechanics (and professional mechanics too) to focus on setting the rocker arm's contact patch on the valve tip, by manipulating the rocker arm's height and the push rod's length, at the expense of other concerns. This may achieve the most rudimentary aspects of rocker adjustment, and it may be convenient, but it cannot possibly result in an ideal adjustment. The most rudimentary aspects of rocker arm adjustment simply keep the operation of the rocker arm within four parameters; (1) the rocker arm should not contact the valve spring retainer when the valve is fully closed, (2) the rocker arm should not contact the push rod when the valve is fully open, (3) the rocker arm slot should never "bottom-out" against the fulcrum, saddle or stud at either extremity of its motion, and (4) the rocker arm tip should never bear down upon an edge of the valve tip; its sweep pattern does not have to be perfectly centered on the valve tip but it should contact the valve tip in the middle half of the valve tip's surface.
As you assemble a cylinder head you can detect rocker geometry and push rod length problems early on by paying attention to the valve stem heights; the valve stem heights should be equal across the cylinder head. If the valve stem heights are unequal, or if one particular valve stem is higher or lower than all the others, you SHALL run into problems.
There are two types of rocker arm designs to consider, the first is the stud mounted, push rod guided type of rocker arm. The height of stud mounted rocker arms is set by the lash adjusting nut (aka the poly lock). Adjusting lash with this type of rocker arm alters the rocker arm's height, and impacts the rocker arm's geometry. In order to maintain consistency in push rod length stud mounted rocker arms are best adjusted mounted on the engine in conjunction with a fixed length push rod.
The other type of rocker arm is the fixed-pedestal mounted type of rocker arm that fastens securely to the cylinder head's rocker arm pedestal. A fixed-pedestal mounted type rocker arm can provide lash adjustment just as easily as the stud mounted variety, by employing a push rod cup style adjuster. The factory rocker arm and the two Yella Terra rocker arms are all of this second type of rocker arm. The high-end T&D and Jesel shaft mount rocker arms are also fixed-pedestal mounted rocker arms. All fixed-pedestal mounted rocker arms are in fact a type of individual shaft mounted rocker arm; they are more stable and contribute fewer rocker arm induced problems to the valve train as long as the saddle/fulcrum is rigid enough. The height of most fixed-pedestal mounted rocker arm is raised by shimming the rocker arm fulcrum/saddle; it is lowered by removing material from the fulcrum/saddle or by removing material from the pedestal cast into the cylinder head. However Yella Terra offers saddles of varying height for their premium YT-6321 rocker arm. This is a very attractive feature of those rocker arms. Increasing valve length also has the same effect as lowering the rocker arm. This type of rocker arm makes it possible to adjust the relationship between the rocker arm tip and the valve tip independent of the push rod, with the cylinder heads sitting on your work bench.
If you are using the factory rocker arms and determine their geometry requires adjustment, a good starting point is to set the height of rocker arm to position the fulcrum’s pedestal approximately in the middle of the rocker arm slot at 50% valve lift. Do not Tufftride the factory rocker arm parts until after the rocker arm geometry has been sorted out.
It is popular to test rocker arm adjustment with the heads assembled on the short block by coloring the valve tips with a felt tip marker, assembling the valve train with the push rods set to zero lash, hand rotating the crankshaft through two revolutions and inspecting the contact patch pattern on the valve tips. As far as I am concerned, the contact patch does not need to be centered on the valve tip, it just needs to stay away from the edges.
Push Rod Length
Sorting out the rocker arm geometry is a prerequisite for determining push rod length. Due to the age of Cleveland series motors, (1) the original manufacturing tolerances can result in dimensional differences, (2) parts have been mixed and matched over the decades, or (3) some parts have already been refurbished once or twice and worked on by many hands of various skill level. For these reasons you may find each cylinder head requires a different push rod length. The actual length of the push rods you shall order for the engine shall be the sum of the length of the longest or shortest “zero lash” push rod plus a small additional amount. This small additional amount added to the length of the push rod establishes the hydraulic tappet adjustment; i.e. the amount you plan to compress the hydraulic tappet plunger.
The factory fixed-pedestal mounted rocker arms, and stud mounted rocker arms require consistency in rocker arm height amongst the all the rocker arms on each cylinder head (both cylinder heads if possible) so that the push rod length required to set all of the rocker arms at zero lash is within a few thousandths of an inch per cylinder head. The length of the longest push rod required to set all of the rocker arms at zero lash shall be the basis for determining what length of push rods to order.
The fixed-pedestal mounted rocker arms equipped with push rod cup adjusters (such as the Yella Terra rocker arms) do not require as much consistency because the adjustable push rod cups will make up the differences. Start with the push rod cup adjusters screwed all the way into the push rod tips and find the rocker arm requiring the shortest push rod to achieve zero lash. The length of this shortest push rod shall be the basis for determining what length of push rods to order. There is a limit to how far you can screw the adjusters out, so keep an eye out for big differences and resolve any problems.
Hydraulic Tappet Adjustment
One rule of thumb for adjusting hydraulic tappets is to compress the plunger 1/2 of the plunger’s available travel; however it is important to measure the travel of the plunger if that’s your plan. The plunger of a modern hydraulic tappet does not compress as much as the plungers did decades ago. The plunger travel of a Crane roller tappet is only 0.062”. The plunger travel of a 1995 Johnson HT900 tappet I have on hand is 0.125”, whereas the plunger travel of a 1970s vintage HT900 is tappet is 0.187”. I am told the plunger travel of a typical modern hydraulic tappet is in the range of 0.060” to 0.080”. Nowadays the recommended range of hydraulic tappet adjustment when using stud mounted rocker arms is 1/8 to 1/2 turn of the adjusting nut beyond zero lash when the engine is hot (adjustable rocker arms are usually mounted on studs with 3/8-24 or 7/16-20 threads). Decades ago the spec for adjusting small block Chevy tappets was one full turn beyond zero lash, which was supposed to set the tappet plunger in the middle of its travel! This means the plunger in Chevy’s tappet had 0.145” of travel.
Some tappets (such as the Morel HLT hydraulic roller tappets) utilize intentionally limited plunger travel as a method to increase the rpm capability of the tappet. This requires adjustable valve train, and push rod length should be determined following the instructions of the tappet manufacturer.
Longer valves are sometimes required (or at least convenient) for solving 3 problems that crop up when installing a higher-lift camshaft. Longer valves (1) increase the distance between a valve spring retainer and the top of the valve guide, they (2) provide the additional height needed for valve springs which have an installed height that is higher than the installed height of the OEM valve spring, and they (3) raise the height of the valve tip which can be a better choice than lowering the rocker arm when adjusting rocker arm geometry.
Manley severe duty stainless steel valves for the 351C are available off the shelf in +0.100” lengths;
FLAT TAPPET CAMSHAFT SECTION
Flat Tappet Camshaft Issues
On occasion a flat tappet camshaft fails prematurely, usually during break-in or soon after a motor is placed in service. This is something we must consider when choosing to use a flat tappet camshaft. Decades ago we installed flat tappet cams in our motors and never thought twice about the possibility of premature failure. When flat tappet cams fail prematurely today there is a logical reason behind the failure. I believe the failures must boil down to one of four conditions:
(1) A lubricant issue (i.e. insufficient wear protection)
(2) A quality control issue (i.e. the parts were not made of the same materials used decades ago or were not surface hardened properly)
(3) A performance issue (i.e. the flat tappet lobes of today’s street cams have faster lift rates and utilize more valve spring force, therefore the cams wear like race cams did decades ago)
(4) Improper break-in (camshaft lobes are splash lubricated, in order to insure adequate lubrication during break-in the motor must be run above 2000 rpm as soon as it is started up)
Considering those 4 possible reasons for premature flat tappet camshaft failure, my strategies to prevent possible failure are:
• Avoid using the highest lift rate camshaft lobes or unreasonable valve spring force. Valve lift no more than about 0.570" theoretical (i.e. about 0.550" factual). Hydraulic intensity no less than "about" 52, i.e. stay away from Comp Cam's Extreme Energy cams and Lunati's VooDoo cams. Major intensity (solid tappet) no less than "about" 44. Valve spring force no more than 130 pounds seated or 330 pounds over the nose.
• Purchase the cam from a trustworthy grinder. There is one manufacturer who is (in my estimation) the source of 95% of all failed valve train parts.
• Custom order the cam requesting the cam grinder's best surface hardening treatment (nitriding) and best lobe polishing.
• Cam cores come in different quality levels, the lobes are narrower in some cases, and lobe taper may vary. A quality cam should have .002" taper on the lobe to aid in lifter rotation and break-in. Yet economical cams may only have about .0005" taper. So when you're custom ordering the cam touch upon the subject of core quality and lobe taper. Make sure you specify the best quality cam core, and specify 0.002" lobe taper.
• Use flat tappets manufactured in North America or Australia with trustworthy quality (Johnson HT900 for instance).
• Use motor oil having very high wear protection properties for both break-in AND normal operation (Valvoline VR1 for instance).
• NEVER use break-in oil because break-in oil has low wear protection properties. Break-in oil is not intended for cam lobes, it is intended to help rings seat, but modern rings and modern cylinder honing techniques preclude the need for break-in oil.
• NEVER use an oil additive; the high zinc ZDDP additives diminish the wear protection properties of a good motor oil.
• Run the motor above 2000 rpm during the entire 30 to 45 minute break-in period of the camshaft to insure the camshaft and tappets are "splash lubricated" adequately.
Distributor Gears for Flat Tappet Camshaft Applications
Iron camshaft cores, such as the cores used for all 351C flat tappet camshafts, are compatible with the original equipment distributor gears found on both factory and aftermarket 351C distributors.
Hydraulic Flat Tappets
The Speed Pro (Johnson) HT-900 hydraulic flat tappet has been a reliable choice for decades. Johnson once boasted of the superior heat treatment of their tappet, they also claimed decades ago when the 351C was a popular motor that their tappet metered oil properly for the 351C. The tappet is sturdy enough for performance usage and sturdy enough for the weight, the valve spring forces, and the canted valve geometry of the 351C 4V valve train. It is also available as an anti-pump-up lifter, part number HT-900R, which requires adjustable valve train.
Valve Springs For Flat Tappet Camshafts
The 351 Cleveland is equipped with a “big block” style valve train composed of large - heavy valves, large-heavy springs and spring retainers, and high ratio rocker arms. According to Crane Cams achieving the best compromise between performance and acceptable valve train wear with a flat tappet “big block” valve train such as this requires setting the valve spring force between 115 to 130 pounds on the seat and no more than 330 pounds over the nose.
The best valve spring for flat tappet street applications I am aware of at this time is Crane Cams #99839, which is a single spring with damper style valve spring. This spring was designed for AMC V8 applications, which is a motor with a “big block” style valve train similar to the 351C valve train.
ROLLER TAPPET CAMSHAFT SECTION
Hydraulic Roller Tappet Valve Train Issues
A Crane hydraulic roller tappet is 44% heavier than a Johnson HT-900 hydraulic flat tappet (148 grams verses 103 grams). Taking into account the Cleveland 1.73:1 rocker arm ratio the heavier roller tappet is predicted to reduce the rev limit of a motor by 650 rpm. Hydraulic roller camshaft lobes also lift valves open at a higher lift rate than the lobes of hydraulic flat tappet camshafts. Lifting a heavier valve train component (i.e. the roller tappet) at a faster rate increases the inertia of that component and makes the roller tappet more likely to lose contact with the camshaft lobe at maximum lift when the camshaft lobe’s nose stops lifting the tappet (objects in motion tend to stay in motion). These are the reasons why hydraulic roller cams can negatively impact the high rpm capabilities of a motor, why hydraulic roller camshaft valve trains have more instability problems, and why they require more valve spring force (both seated and at maximum lift) to maintain valve train stability.
The roller and roller axle of a hydraulic roller tappet are splash lubricated as opposed to pressure lubricated. The amount of splash lubrication occurring at idle or low engine speeds is insufficient for the roller and axle to sustain heavy loading, therefore the wear rate of those parts increases as spring forces increase. Although a hydraulic roller cam valve train requires additional spring force to maintain valve train control and stability, the amount of spring force that can be applied is not limitless. With a high-lift hydraulic roller cam we walk a line between applying sufficient valve spring force for good control of the valve train yet keeping that force light enough for acceptable roller tappet wear.
As a camshaft lobe turns beneath a tappet, lifting the tappet at the same time, it imparts a side thrust force against the tappet, in effect trying to push the tappet against the wall of the tappet bore. This is due to the fact that the ramps and flanks of a camshaft lobe are ground at angles; a lobe does not contact a tappet in such a manner as to push it perfectly upward inside the tappet bore. The angle of the side thrust (and therefore the strength or magnitude of the side thrust) acting upon the tappet is dependent upon the radius of that part of the tappet that contacts the camshaft lobe. A flat tappet’s face is ground on a 50” radius, whereas the roller of a roller tappet has about a 0.35” radius. This is a critical difference between flat tappets and roller tappets. A flat tappet has very little side thrust acting upon it because the angle of that thrust is practically parallel to the axis of the tappet bore. A roller tappet on the other hand has a significant amount of side thrust acting upon it pushing the tappet directly against the tappet bore, making the roller tappet’s body prone to distortion. The internal parts of a hydraulic tappet are some of the most precision manufactured parts in the entire motor, the clearances are critical, the tappet cannot function properly if the body distorts. Thus it is critical that the body of a high performance roller tappet is made sturdy enough to prevent its distortion even when subjected to higher valve spring forces and higher engine speeds.
Distributor Gears for Roller Camshaft Applications
Steel camshaft cores, such as the cores used for roller cams ground by Crane Cams and Bullet Racing Cams, require a compatible steel distributor gear. Crane Cams manufactures the steel roller cam cores used by all the cam grinders, and they also manufacture the proper steel distributor gear for use with camshafts ground on their cores. Crane #52970-1 is the gear for 0.500” distributor shafts; Crane #52971-1 is the gear for 0.531” distributor shafts. The gear for 0.531” shafts is also available via Ford Racing Performance Parts under part number M-12390-J.
Hydraulic Roller Tappets
Although the pricing is tempting the Ford factory 5.0 HO hydraulic roller tappet is not recommended for use in your 351C. The Ford tappet has been problematic in 351C applications. There are four reasons for this: (1)The 351C valve train is heavier than the valve train the 5.0 tappet was designed for; (2) the 351C valve train utilizes higher valve spring forces than the 5.0 tappet was designed for; (3) the 351C valve train geometry and splayed push rods subject the tappet to side thrust forces greater than the forces the 5.0 tappet was designed for; and (4) the waist machined into center of the 5.0 HO tappet is too high, it has been found to rise above the top of the lifter bore at maximum lift and dump the engine’s oil pressure in some 351C blocks.
The aftermarket hydraulic roller tappets sold by Crane Cams and Morel are known to operate reliably in Cleveland applications. The waists machined into the center of these tappets do not rise above the top of the lifter bore at maximum lift, and the tappet bodies are thicker and therefore resist distortion (with the penalty of increased weight).
Crane Cams manufactures one hydraulic roller tappet for 351C applications, part number 36532-16, and it’s a good one. Crane’s tie-bar style roller tappet is machined from 8620 steel billet and it is heat treated. A precision fit plunger assembly is used to provide the proper bleed-down rate, permitting high RPM use in properly set-up engines. The strength of the heat treated 8620 material prevents distortion of the lifter body, thus permitting more consistent operation in high spring pressure and in high RPM applications, due to the consistency of the plunger to tappet body clearance. Crane hydraulic roller tappets weigh 148 grams (that’s half the weight of a pair). Internet pricing for a set of Crane’s tappets is in the range of $635. This is my preferred hydraulic roller tappet.
Morel does not sell their tappets directly to the consumer; their tappets are sold via a network of retail businesses several of them being cam grinders, including Lunati in the US and Crow in Australia. I have not been able to verify the weight of Morel’s tappets. Morel manufactures three 0.875” OD hydraulic roller tappets for 351C applications:
(1) Morel hydraulic roller tapper #5323. This tie-bar style roller tappet is described as a “street” tappet with an upper rpm limit in the range of 6200 rpm to 6500 rpm. Internet pricing for a set of these tappets is in the range of $380.
(2) Morel hydraulic roller tappet #5327. This tie-bar style roller tappet is described as a hydraulic-limited travel tappet (i.e. HLT). I assume this means they are designed for higher rpm and that they require adjustable valve train. Internet pricing for a set of these tappets is in the range of $505.
(3) Morel hydraulic roller tappet #5879. This tie-bar style roller tappet is described as a “pro” high rpm HLT tappet. It is designed for oil viscosity no greater than 5W/40. Since it is a “HLT” style tappet I assume it requires adjustable valve train. Internet pricing for a set of these tappets is in the range of $830.
Valve Springs For Hydraulic Roller Tappet Camshafts
PAC Racing Springs is a small division of the Peterson American Company (i.e. PAC) the largest spring manufacturer in the USA. They manufacture the ovate wire beehive valve springs that have become so popular in the performance industry. The ovate wire beehive valve springs are manufactured in two series, the 1200 series and the 1500 series. The 1200 series valve springs are the budget springs. The 1500 series valve springs are nitrided, polished and nano-peened; they are easily identified by their GOLD COLOR. If your bee hive valve springs are not gold colored, they are not the springs I am recommending, and you must live with the consequences of YOUR choice. The 1500 series valve springs cost about 25% more than the 1200 series valve springs, they are the springs I recommend. There are also low priced substitute beehive springs on the market … buyer beware.
The #1520 Big Block Chevy beehive spring manufactured by PAC Racing Springs is a good choice for 351C hydraulic roller cam applications, since the Big Block Chevy’s valve train is very similar to the 351C valve train.
PART 8 - PREPARING A 351C RACING ENGINE
The production 351C was never intended for high rpm racing (8000+ rpm) but that didn't stop people from doing so. When the production engine is set-up for racing (excepting the connecting rods) it will withstand those sort of engine speeds and higher for a while before something breaks. The 351C was duty-cycle-tested up to 7000 rpm which was a high rpm duty cycle for an engine intended for mass production circa 1968. Based on my experience, and being conservative, I'd say the production block, crank, connecting rods and cylinder heads are good for many years of racing if engine speed is limited to about 7000 rpm; even the 2 bolt main caps resist "walking" at 7000 rpm! However, for reasons I shall explain below I don't recommend spending money to prepare the production connecting rods for racing unless the rules require using them. There are two caveats regarding the production engine block: the lubrication system and the thin cylinder walls require steps taken to amend their shortcomings. Another consideration, any partially counter-weighted crankshaft that has been designed for maximum bob-weight instead of minimum bearing load increases the loading of the second and fourth main bearings and bulkheads, and cracking of those bulkheads is a possibility. In the end, the durability of a 351C racing motor shall hinge upon the supporting parts that are selected and the time, money and detail invested in preparing it. I don’t claim to be an expert. But if you may find what I've learned over the decades helpful, here's the synopsis. The following set-up info is good for all types of competition excepting drag racing.
I'd prefer to build a racing motor around a heavy duty block designed for that purpose; for instance the US manufactured racing block known as either the SK block or the 366 block, the Australian manufactured XE192540 NASCAR block or the new Tod Buttermore block. The heavy duty block would be a more durable choice, having thicker cylinder walls and thicker bulkheads. These blocks are less likely to fail during the abuse of racing, therefore they make a good insurance policy against wasting the money you've invested in preparing the race engine. The price of replacing racing parts is expensive, as is the price of machine work and the price of assembling a racing engine. If one or two production blocks fail over the course of several racing seasons then using a production block would end up costing more money in the long run. Thicker cylinder walls make it possible to use more compression and higher engine speeds. Sturdier bulkheads make the block more compatible with a less expensive partially counter-weighted crankshaft such as the factory crank or a "sportsman" crank ... although I would still prefer to use a fully counter-weighted crankshaft if it is in the budget.
If I planned to use the iron 4V heads then the block material would be iron as well, whereas aluminum heads can be mated to an iron block or an aluminum block. Besides the weight reduction additional horsepower can usually be coaxed from aluminum high-port racing heads (not because they are made of aluminum, but because they have higher ports and possibly high swirl combustion chambers), but there is a limited selection of intake manifolds available for the high-port heads and they require custom manufactured exhaust headers in many applications. Iron racing blocks are both sturdier and less expensive than aluminum blocks. However a heavy duty aluminum block from Tod Buttermore and a set of aluminum heads shall reduce the weight of a race car by a significant 200 pounds (91 kilograms). Of course the benefits of weight reduction must be weighed against the higher price and the lesser durability of the aluminum block. Some guys argue in favor of an aluminum block by pointing out it is often repairable when damaged whereas an iron block is not.
If I intended to use the production block, I would accept the compression ratio and engine speed limitations inherent in that choice. The De Tomaso factory determined circa 1973 that to avoid failure of their racing engines employing the production block they had to limit those engines to 7000 rpm and 10.5:1 static compression. Operating within those limits the engines produced about 500 horsepower. There are choices however in parts and machine work that shall reduce the possibility of failure or raise the operating limits of the production block.
In terms of preparing the block for racing I’d sonic check the cylinder walls to establish their thicknesses; insuring the walls are at least 0.120” thick on the thrust sides after boring and at least 0.080” thick on the non-thrust sides after boring. I'd have the crankshaft main bearing saddles align honed. I'd level the block's decks, setting the decks up with a finish compatible with multi-layer steel (MLS) head gaskets. I'd index the boring machine to the crankshaft's axis during the boring process to insure the cylinders are perpendicular to the axis of the crankshaft. A piston trying to stroke up and down in a cylinder that is canted to the front or rear must operate in a “wedged” manner that puts an abnormal load on the cylinder walls and causes floating wrist pins to hammer out their locks. A piston will operate in a cocked manner if a cylinder is canted to the left or right which again puts an abnormal load on the cylinder wall. This abnormal cylinder wall loading contributes to cylinder wall cracking, therefore indexing the boring machine to the crankshaft's axis helps to alleviate cracking of the production block's thin cylinder walls. It also reduces frictional losses and makes more horsepower! The cylinders would also be bored and honed with head plates and main bearing caps torqued in place for the best possible ring seal, which also makes more horsepower. I'd install lifter bore bushings in all 16 lifter bores if the block incorporated the factory 351C lubrication passages (the Buttermore block incorporates a main priority lubrication system). The bushings would have 0.060" orifices. I'd install cam bearing oil passage restrictors at all 5 cam bearings no matter what block I'm using, also with 0.060" orifices. I would use MLS head gaskets for racing. The production block’s rear main seal is a rope seal; I’d replace it with a neoprene seal which requires pulling a small pin from the seal groove in the rear main bearing cap and filling the pin hole with a dab of sealant. The main bearing caps and the heads would be clamped with studs instead of the factory bolts. Limiting the production block to a static compression ratio in the range of 10.5:1 to 11.0:1 (8.0:1 dynamic compression) is another measure that can be taken to prevent cylinder wall cracking. I know guys who would scoff at the idea of racing with the production block set at 11.0:1 static compression, but I'm not sure if they installed full round skirt pistons in their race motors. Building the motor around a racing block having thicker cylinder walls provides more latitude to set the compression ratio higher, if the fuel the motor shall be operating on and the cylinder head combustion chamber design allowed it. I'd prefer to lubricate the motor with a dry sump lubrication system because a wet sump system is not ideal for coping with the g-forces encountered while cornering, accelerating and braking on modern race tires. However, if I intended to use a wet sump lubrication system then I'd plan to use a high capacity oil pan which incorporates baffles with hinged doors, a windage tray and a scraper. The wet sump system would also incorporate a high capacity oil accumulator (i.e. an Accusump). Regardless if the car is equipped with a dry sump lubrication system or a wet sump lubrication system it MUST be equipped with an oil cooler.
The aftermarket stroker crankshafts are manufactured as inexpensively as possible using Chinese castings or forgings, the quality of their machine work is adequate at best (and often inadequate), and their quality control is poor. I would not consider using a crankshaft manufactured to such standards for a street engine OR a racing engine. Nor would I use a crankshaft with more than a 3.50" stroke, the additional crank-arm leverage & piston speed is not beneficial in terms of sports car racing, road racing, track racing or circuit racing. In regards to selecting a crankshaft for a racing engine, there are three viable choices. Choice number 1: The first choice is to utilize the production nodular iron crankshaft. The production crank has a track record of quality and durability. Since the factory crank was externally balanced I would have it internally balanced, which increases the price of using the factory crank. Since the production crankshaft was designed for maximum bob-weight as opposed to being designed for minimum bearing load it heavily "loads" the second and fourth main bearings and bulkheads during ultra-high rpm operation. This means I'd choose to limit engine rpm to about 7000 rpm. Choice number 2: The second choice is to purchase a mid-price "sportsman" style forged steel crankshaft. A forged steel crank should be tougher than a cast iron crank, but realistically we must keep in mind they are based upon Chinese forgings. The good ones are machined in the US by reputable companies. The sportsman crank should come from the manufacturer internally balanced, however like the production crankshaft a sportsman crank is only partially counter-weighted. So before you purchase one shop around, talk to the engineers who designed them, and make sure you purchase one that has been designed to minimize bearing loads. Otherwise it shall heavily load the second and fourth main bearings and bulkheads during ultra-high rpm operation just like the factory crankshaft. Some crankshafts also offer improved rod bearing lubrication passages, which is another topic to discuss with the engineers before you make your choice. Choice number 3: The third choice is to purchase a very expensive, fully counter-weighted, steel crankshaft (forged or billet). The fully counter-weighted crank is best at reducing the "loading" of the second and fourth main bearings and bulkheads during ultra-high rpm operation. The bending deflection across the center main at high loadings and high engine speeds causes measurable power losses in engines equipped with partially counter-weighted crankshafts. Therefore the benefits of a fully counter-weighted crankshaft are less stress on the engine block and the reduction of power losses, i.e. an increase in power output!
Regardless of which crankshaft choice I make, I'd have the crankshaft magnafluxed, tufftrided, polished, and dynamically balanced (remember the production crank should also be internally balanced). If the motor is set up for 10W, 15W or 20W oil then I'd use fully grooved copper-lead alloy main bearings (Mahle/Clevite MS-1010P). Acquiring fully grooved bearings requires using the upper halves of two sets of standard bearings. If the motor is set up for 0W or 5W oil then I'd use 3/4 grooved copper-lead alloy main bearings (King Bearing # MB5169HP). The main bearings would be set-up with 0.0009" to 0.0011" clearance per inch of main bearing journal diameter which is how they were set-up 40 years ago and is still fairly common these days. Using the factory crank with a heavy piston & rod combination will require more rod bearing clearance than what is customary however (0.0011" to 0.0013" clearance per inch of rod bearing journal diameter). I'd use the ATI #918920 neutral balanced steel crankshaft damper. This damper has a reputation for preventing cracking of the second and fourth bulkheads when the factory crankshaft is used. I prefer the durability of a light weight neutral balanced steel flywheel (Yella Terra YT9902N) over the additional weight-loss of an aluminum flywheel.
I'd use piston and rod assemblies with floating pins, my preference being to use 6.00" long connecting rods. The 6.00" rods are gentler on the cylinder walls, they are gentler on the piston skirts, they rock the pistons less in the bores and since they require "shorter" pistons the weight of the pistons is reduced. As long as the rod length to stroke ratio does not exceed 1.72:1 the longer rods will not impair acceleration or create induction system "lag" issues. The 6.00" rods are actually small block Chevy rods; using such rods requires a crankshaft with 2.100" rod journals in order to compliment Chevy diameter rod bearings. You'll find that is the standard journal diameter for aftermarket crankshafts. The standard big-end width of a Chevy connecting rod is 0.940" however. Chevy rods in which the big-ends had been narrowed to 0.831" in order compliment Ford width rod journals (intended for use with Mahle/Clevite CB1227 rod bearings) were once readily available, but that no longer seems to be the case. You may have to order custom rods for this application, but don't let that deter you because there are several excellent choices in reasonably priced custom manufactured rods on the market; the sport rods from Howards Racing Cams are an example.
It has been customary to turn-down the rod journals of the factory crankshaft to 2.100" when using it with 6" connecting rods, but if you must custom order the 6" rods why not specify having the rods machined for standard 351C size rod journals and standard 351C rod bearings? Another option in a 6" long connecting rod (actually 5.956") that is compatible with the standard 351C rod journal diameter (2.31") is the Eagle H-beam connecting rod for the 351W, #CRS5956F3D. The crankshaft’s rod bearing journals do not require being re-ground because the 351W has the same size rod journals as a 351C (2.311”). The 351W uses a Clevite CB831 rod bearing however, which has a different shell thickness but the same width and ID as a 351C bearing. The 351W rod also uses the same size wrist pin as a 351C (0.912”).
However, for those preferring to use 5.78" production length connecting rods (and therefore production 2.31" diameter rod journals) it doesn't make financial sense to use the factory rods for racing unless the rules require them. The factory rods require magna-fluxing, shot-peening, 180,000 psi rod bolts, re-sizing the big ends and installing bushings in the small ends for floating wrist pins. Thus prepared the factory rods still lack locating dowels for the big end caps, and are only reliable to about 7200 rpm. The price difference between setting up a set of production rods in that manner and purchasing a set of Eagle #CRS5780F3D H-beam rods is almost nil, yet the Eagle rods are better quality rods, and are reliable at higher rpm (the Eagle rods use Mahle/Clevite #CB831 351W bearings).
Regardless of the length of the connecting rods, they should be used in conjunction with full round skirt forged flat-top endurance racing pistons. The combination of a 6" rod and a round skirt piston has an excellent track record for preventing cracking of the production block's thin cylinder walls. The Ross pistons are currently available for 4.030" bores in pin-heights for factory length connecting rods or 6" long connecting rods from Summit Racing at a very good price. The Ross #80556 pistons with 1.668" pin height are for production length rods, the Ross #80566 pistons with 1.446" pin height are for 6" long rods; the second pistons also use Chevy diameter wrist pins. Using a 351W rod will require a custom round skirt piston having a pin height in the range of 1.47” to 1.495”.
Assuming the cylinder heads are designed to use 351C valve train parts, I would use Yella Terra YT6321 or T&D Machine #7200 or #7201 rocker arms. It is common these days to employ 1.8:1 ratio rocker arms for the intake valves and 1.7:1 or 1.6:1 ratio rocker arms for the exhaust valves; the T&D rocker arms are available in several rocker arm ratios. On the intake side I'd use Manley's #11872-8 light weight race master 4V stainless steel intake valve with a titanium spring retainer. Of course, if I wanted to maximize the life of the valve train I'd opt for titanium intake valves which would allow me to select softer valve springs or rev the motor to higher rpm. On the exhaust side I'd use Manley's #11805-8 severe duty 4V stainless steel exhaust valve with a chromoly spring retainer. One caveat here, when juggling the weight of the valves its safer to set-up the valve train so the intake valves are the first to float, because it’s usually the exhaust valves that hit the pistons first so you want to avoid floating the exhaust valves. I would not bolt the cylinder heads on the motor until the rocker arm geometry is set-up properly (see my notes on this in the valve train section above). The pedestal mounted rocker arms I've recommended make this possible. I'd operate the valve train with 3/8" OD push rods made from 0.080" wall thickness seamless chromoly tubing. I'd select a camshaft with lobes that are as conservative as possible in terms of ramp design and lift rate, while keeping the motor competitive. It is important to realize that not all camshafts are created equal; some lobes are tougher on the valve train than others. PAC Racing valve springs would be selected to complement the cam, tappets and the application.
I'd set the rev limiter of a race engine using the production block somewhere around 7000 to 7200 rpm. A heavy duty block employing a fully counter-weighted crankshaft can rev much higher. Rock-n-Roll!
Suppose I just want to freshen up/upgrade my current 351C...maybe 400hp?
Is this possible and could I then continue UP to 500 if 400 isn't enough?
I don't see much difference in building a motor for 400 BHP verses building one for 500 BHP, I wouldn't recommend doing anything differently. Engine speed is the real issue, and what I've described is how to build a tough motor that isn't going to break when used for high performance street and sports car applications, with a red-line in the range of 6500 to 7000 rpm.
If what you are suggesting is hot-rodding the motor with limited preparation, some guys get away with it, and some don't. Some of that depends upon just how hard the motor will actually be run after its hot-rodded. Some of that depends upon what shape the motor is in right now. I don't know how hard you or the other people reading this will run your motors, so I have to be careful and thorough, and make recommendations that won't result in damage to you, your motor, or your car. I have to assume that at the very minimum you'll run the motor on a high rpm blast every once in a while, like accelerating up a freeway on-ramp, or when you drop a gear or two to pass somebody on the freeway. I have to assume you may succumb to temptation on a lonely stretch of road and wind the motor out WFO to see what 150 MPH feels like.
I have "freshened up" 351C's before, even back in the '70s when the motors were still relatively new. But as I "freshen up" the motor I also toughen it up to prepare it for high performance. As it comes from the factory the 351C 4V (M code) isn't prepared for engine speeds much higher than 5000 rpm, the 351 Cobra Jet (Q code) maybe 6000 rpm, but not 6000 rpm for extended periods of time. The 351C 4V is not at all prepared for the kind of hard flogging owners have given it over the years. The ruggedness is there in the major castings, its the small details that need improving; lets put it this way, it needs to be brought up to Boss 351 spec.
To have confidence in the durability of a 351C 4V on a track day, or powering a Pantera or Mustang being flogged on a winding mountain road at high speed; we need to toughen the motor up while we freshen it up. Here's how I would approach freshening up the motor, the choices I would make.
First, here's a link to a thread with another post of mine with tons of very well "edited" and up to date rebuilding information: Cliff Notes Version of Sticky #3
The 351C 4V has 5 weaknesses that stick out in my mind.
(1) Over the last 4 decades many 351C owners have experienced the misfortune of having one of their 4V valves drop its brittle head while the motor is running, resulting in devastating damage to the motor. The valve head breaks off just below where it is welded to the stem. It has happened to completely stock motors, mildly modified motors and extensively modified motors too. The damage can occur while the motor is idling in the driveway, cruising on the road or wicked fully open (WFO). There is no way to predict if or when it will happen. The fact that a motor has run for 4 decades using the original valves is no guarantee it wont drop a valve tomorrow. If your 351C 4V is still equipped with the original Ford valves, it is a ticking time bomb.
The 4V valves are also designed with 4 groove, loose fitting valve spring locks, which are universally shunned for high performance use, higher valve spring force, higher rpm. So that's the second reason for replacing the valves with single groove valves.
My preference is Manley valves:
*Manley Performance severe duty intake valve, 2.08" dia. x 5.24” long (std. lg.); #11762-8, 131 grams
*Manley Performance severe duty 2V exhaust valve, standard length, #11807-8, 102 grams
4V intake valves
* Manley Performance severe duty 4V intake valve, standard length, #11800-8, 139 grams
(8 grams lighter than the factory intake valve)
* Manley Performance race master 4V intake valve, standard length, #11872-8, 129 grams
(18 grams lighter than the factory intake valve)
4V exhaust valve
* Manley Performance severe duty 4V exhaust valve, standard length, #11805-8, 108 grams
(15 grams lighter than the factory exhaust valve)
Consider this, by using the lightweight Manley Race Master intake valves in conjunction with titanium valve spring retainers you can appreciably lighten the intake valve train, improve the valve train dynamics, and raise the rev limit of the motor. You can save money and use chromoly valve spring retainers in conjunction with Manley Severe Duty exhaust valves.
Whatever brand of valves anyone chooses, it is imperative the stainless steel valves have hardened steel tips. If the heads need seat inserts cast iron (or beryllium-copper) valve seats are complimentary to stainless steel valves. To prevent rapid wear of stainless steel valves in the valve seat area the cylinder head’s intake seat width should not be less than 0.060” and the exhaust seat width should not be less than 0.080”; seat run-out should be 0.001” or less. I would equip the cylinder heads with silicon-bronze valve guides to best compliment stainless steel valve stems. I would utilize spring loaded elastomer valve stem seals such as Ford Racing Performance Parts #M-6571-A50 or Manley Performance #24045-8; installation of this type of seal requires machining of the top of the valve guide to 0.530” diameter.
(2) The production connecting rod nuts were the weakest link in the 351C reciprocating assembly. The threads will strip out of the nut when the motor is run hard and the result is major carnage.
An easy upgrade is to simply replace the nuts with parts from ARP, #300-8371. The connecting rod bolt can be re-used. If you replace the bolts the big-ends of the rods will have to be re-sized. Leaving the bolts in place and replacing only the nuts will avoid having to have the rods re-sized. The nuts truly are the problem
(3) The ring of the OEM crankshaft damper is not bonded to the hub, and it is too light for performance use (with the exception of the dampener found on the Boss 351). PLUS the OEM dampeners are decades old today, it is a common malady of all unbounded crankshaft dampers that as they age the ring spins on the hub or slowly creeps forward or backward off the hub, therefore the dampener MUST be replaced. Purchasing a new R code dampener is not possible, so a good choice for a frugal engine project is a 100% steel SFI approved Romac #0203 dampener.
(4) The 351C lubrication system has a propensity for low oil pressure and worn bearings. The basic design flaws of the lubrication system are that there's no control of where the oil is flowing nor is there control of how much oil is flowing. The large port in the walls of the tappet bores is another design flaw. It allows too much oil to flow to waste, it creates tappet compatibility issues, and it allows cavitation resulting from the motion of the tappets to spread within the oil passages. The lubrication system problems impact solid tappet motors and hydraulic tappet motors equally. The symptoms are the same regardless if the rev limit is 5000 rpm, 6000 rpm, 7000 rpm or higher; the symptoms merely worsen as rpm increases. If I am not going to disassemble the motor it shall not be possible to fix the lubrication system right (install tappet bore bushings or cam bearing restrictors).
The bottom line in regards to an unmodified lubrication system, the motor should have more than 50 psi hot oil pressure (the target is 60 psi) from about 2000 rpm all the way to the rev-limit. The oil pressure is low because the volume of the oil pump is over-taxed. The motor will have 60 psi hot oil pressure with the standard volume oil pump and the standard oil pump spring as long as the capacity of the oil pump is not over-taxed. Therefore the proper way to boost hot oil pressure (and my first step) is to limit excessive oil leakage and limit the amount of oil flowing to the cam bearings or valve train. About the only thing we can do without disassembling the motor is to limit the amount of oil flowing to the valve train. Assuming we're installing a hydraulic cam, I would purchase a set of 5/16" push rods with 0.120" thick walls, these have a 0.072" passage over 8" long down the middle. Or 3/8" pushrods, with 0.080" walls, with 0.040" restrictors in the tips. The restricted push rods are a decades old recognized method to control the amount of oil flowing to the valve train with hydraulic tappets.
Limiting the amount of oil flowing to the valve train will help, but it will most likely not be adequate to boost the hot oil pressure to 60 psi, maybe not even above 50 psi. So the question becomes what else should we do to "boost" the pressure. I would not use motor oil thicker than 20W50, that only slows down the already inadequate amount of oil flowing to the reciprocating assembly. I've used the Moroso #22850 high pressure oil pump spring in the past ... but I've changed my mind about that. A standard volume oil pump with a 1/8" to 3/16" thick washer between the oil pump spring and the pin which retains it may raise oil pressure enough. A high volume pump will boost pressure too, and it is probably the best measure to take, but there are drawbacks to that. I'm told the standard volume oil pump cavitates, therefore the HV pump will cavitate worse. I've also seen high volume pumps turn motors with a lot of miles on them into serious oil burners because they throw more oil on the cylinder walls. If neither of those "tricks" will boost hot oil pressure above 50 psi then the motor has an excessive oil leakage problem and it should be disassembled.
(5) Cylinder walls that crack is a weaknesses of the production block.
Part of freshening up a motor often includes a new set of rings. If I decide to replace the rings, there's an opportunity to replace the pistons too, I would purchase a set of forged full round skirt flat top pistons, i.e. endurance racing pistons. Summit Racing sells the Ross #80556 round skirt piston for about $586 a set.
These are twice the price of TRW flat tops, but that is a great price for the Ross pistons which used to cost over $800. The round skirt pistons are a proven way to prevent cracking cylinder walls. If I'm going to buy pistons, I prefer to spend the extra $280 for the durability it provides the motor.
Unplanned things can happen during a "freshening-up". Machine shop work tends to snow-ball.
For instance, any freshening up involves a valve job. In this case, I'm having the valves replaced too. If the heads need seat inserts I'll specify cast iron valve seats which are complimentary to stainless steel valves. To prevent rapid wear of stainless steel valves in the valve seat area I'll specify the intake seat widths should not be less than 0.060” and the exhaust seat widths should not be less than 0.080”; seat run-out should be 0.001” or less. I'll also specify a 3 angle valve job. I would equip the cylinder heads with silicon-bronze valve guides to best compliment stainless steel valve stems and I'll utilize spring loaded elastomer valve stem seals such as Ford Racing Performance Parts #M-6571-A50 or Manley Performance #24045-8; those seals require machining of the top of the valve guide to 0.530” diameter.
Another scenario, I pull the heads & discover the cylinders look pretty bad. The amount of horsepower a motor will be capable of making is very dependent upon how well the rings seal in their bores, so I can't be sure how much horsepower it shall achieve "for the time being" if the cylinders & rings aren't put into good shape. And if I am going to have the cylinders bored, I would have them indexed to the crank during that procedure.
I prefer to index the cylinders to the crankshaft "WHEN" the cylinders are bored because if a cylinder is canted to the left or right it will cause the piston to cock in its bore, placing extra load on the cylinder wall, that would make the cylinder wall more prone to cracking. So just like round skirt pistons, indexing is a way to help the cylinders walls resist cracking. If a cylinder is canted to the front or rear it will cause floating wrist pins to hammer out their locks; so indexing is a prerequisite for floating pins too. If you think you may ever want floating wrist pins, that is another reason to have the cylinders indexed to the crank when you have them bored. The 351C doesn't have enough cylinder wall thickness to bore repeatedly, you can't put-off indexing them until the next time.
The prelude to indexing the cylinders is to have the bearing saddles align honed, and the machine shops will want to level the decks at the same time too. So, if you're gonna touch the cylinder walls with anything more than a ridge reamer and a glaze breaker, you may as well plan on going the full route with block machining, its unavoidable.
If a machinist is going to service the block, I would plan to modify the lubrication system at the same time with the simple installation of 5 cam bearing restrictions and 16 tappet bore bushings (having 0.060" orifices). I can do this myself at home with about $420 worth of parts. This flat out fixes the lubrication system once and for all and makes it perform admirably. People will comment its not needed for a street motor, but after fooling around with these engines as long as I have I disagree. Besides, its so inexpensive to fix the lubrication system, I believe its illogical not to fix it.
The crankshaft must be rebalanced whenever parts (like pistons) that weigh more or less than the original parts are substituted. So, if I'm replacing pistons, the weight of the Ross pistons will have to be compared to the weight of the OEM pistons, and if they are heavier we may be in luck, and can adjust them to weigh the same as the OEM pistons. If not, then the crank will need re-balancing, and the machine shop is going to want to regrind the bearing journals before they balance the crank. The journals of a nodular iron crankshaft must be polished after they've been reground too.
If I arrive at this situation I would go the full route towards achieving durability. I would have the crankshaft ground to these clearances (0.0020" - 0.0025" mains; 0.0025" - 0.0030" rods). Then I would have it tufftrided (surface hardened), straightened, and micro-polished. I would purchase a lightweight steel flywheel (manual trans motors). And finally I would have the entire reciprocating assembly dynamically balanced before reassembly. Dynamic balancing takes a lot of vibration out of the motor, it makes it run smoothly. This accomplishes 3 things: (1) it makes the motor feel much more high quality (2) it makes the motor more inviting to rev at high rpm (3) it makes the motor more durable.
The crankshaft rear main seal is a rope seal, its usual to replace it with a neoprene seal, this requires pulling a small pin from the seal groove in the rear main bearing cap and filling the pin hole with a dab of sealant.
Back to the subject of a freshening up ...
The single most important thing to do to improve the performance of the engine is to raise the static compression ratio to achieve 7.6:1 dynamic compression with whatever camshaft you choose to use. Don't stick a hotter cam in a low compression 351C!
The factory connecting rods are fine for street motors and an occasional blast to 7000 rpm ... as long as the rod nuts are replaced with the ARP rod nuts. Pressed pins are just fine for the street too.
The pin height of the Ross pistons will raise the static compression ratio a about 3/10ths. To compute the dynamic compression ratio I need to decide what cam I'm going to use, so I'll know when the intake valve closes. Then I can juggle the static compression ratio to obtain a suitable dynamic compression ratio (about 7.6:1 dynamic c/r is good for 91 octane pump gas). Once I have a target static compression ratio, I can decide how to achieve it. Milling cylinder heads is my preferred way to adjust compression ratio, rather than taking a bunch of material off the decks of the block.
The factory main and rod bearing clearances are too tight for performance usage, and the factory bearings are too soft. When I drop the oil pan I'll likely find ribbons of babbit laying in the bottom.
If the motor is disassembled far enough to replace pistons and connecting rod nuts, then its also disassembled far enough to replace the main & rod bearings. I use Clevite 77 bearings because they are a tougher bearing. They are also available in "plus" sizes, which can help achieve the right bearing clearances even without having the crankshaft re-ground. Those clearances are 0.0020" - 0.0025" mains; 0.0025" - 0.0030" rods. Although it doubles the price of the main bearings I like to use the uppers from two sets of main bearings so that the main journal bearings are fully grooved, 360 degrees. The Boss 351 had fully grooved main bearings, the purpose is to supply lubrication to the rod bearings through 360 degrees of crankshaft rotation, the same reason other performance motors had cross drilled crankshafts or crankshafts with grooved main bearing journals. This doubles the oil flow to the rod bearings. 351C bearing sets used to come fully grooved, but that was discontinued as the viscosity of motor oil was reduced. Fully grooved mains are no problem for 20W50, 15W40, 10W40 or 10W30.
Use a new standard volume oil pump if the motor has tappet bore bushings. Without tappet bore bushings I would use a high volume oil pump OR shim the standard volume pump's oil pressure control spring. I would use a standard oil pump drive shaft (intermediate shaft) with either pump.
I would purchase a good oil pan, because all the preparation in the world won't help the motor if the suction of the oil pump goes dry.
If the motor is still equipped with a breaker point ignition I would replace it with a breakerless ignition; Ford Duraspark, MSD, etc. 20° centrifugal advance in by 3000 rpm, 16° to 18° initial advance, vacuum advance limited to 10° and connected to ported vacuum.
Finally I'll touch on the valve train.
I would replace the camshaft timing set with a new full roller, steel, multi-index (9 keyway) timing set.
Replacement push rods that will not flex (store energy) are always a part of preparing a motor for performance. Push rods are the weakest link of an OHV valve train. Don't be concerned with the weight of the push rods, flexing and harmonics are the issues. The use of premium metals, large OD, thick walls and tapered walls are the ways to combat flexing and harmonics. I like to use 5/16" diameter chromoly push rods with 0.120" thick walls.
The motor needs valve springs compatible with the cam, the red line rpm, and the weight of the valves. If the motor has a hydraulic flat tappet cam acceptable valve spring pressures should fall into the range of 115 - 130 lbs on the seat and 300 - 330 lbs over the nose (roller cams require more spring force than that). Assuming we're using a cam with 0.540" to 0.570" valve lift, that would call for a spring with 350 lbs/inch spring rate that will allow 0.570" lift without binding. That's not much more over the nose spring pressure than what was stock on a Boss 351, but that much valve spring combined with lightweight valves will enable the motor to rev to 7000 rpm without valve float. It is always better to use too much spring force than not enough; always let the cam grinder be the final word on how much spring to use, not me.
The stock rocker arms are OK to 0.615" valve lift if the 5/16" rocker arm fasteners are upgraded with ARP parts(four packs of #641-1500 bolts and two packs of #200-8587 washers) ... and if the geometry is set correctly.
My preference is Yella Terra rocker arms for any and every 351C that is having the rocker arms upgraded. They are a much better design than push rod guided rocker arms which use studs, guide plates and hardened push rods. There are versions that don't require pedestal machining (YT6015) but they are difficult to find in the US. The oem rocker arm pedestals clamp the rocker fulcrum to the head with a 5/16" cap screw, that cap screw limits the use of the oem rocker mounting system to about 7000 rpm and no more valve spring force than 400 pounds over the nose "IF" high strength 5/16" fasteners are used. For higher rpm and/or higher valve spring force the pedestals should be milled and tapped for 7/16" cap screws (to allow the use of the heavy duty YT6321 rocker arms or T&D Machine rocker arms).
To toughen up the 351C 4V while you freshen it up:
(1) Replace the valves (Manley 4V valves: 11872-8 intake & 11805-8 exhaust). These valves use single groove spring locks.
(2) Iron valve seats, bronze valve guides, 3 angle valve job, spring loaded elastomer seals (FRPP M-6571-A50 or Manley 24045-8)
(3A) Valve springs for flat tappet applications rated at 115/130 seated; 300/330 over the nose (such as Crane’s #99839).
(3B) Valve springs for roller tappet applications rated at 150 seated; 370 over the nose.
(4) Steel valve spring cups, titanium spring retainers for the intake valves, and chromoly spring retainers for the exhaust valves.
(5) Push rods made from seamless chromoly tubing, 5/16 diameter X 0.120" wall, or 3/8 diameter x 0.080" wall
(6A) Factory rocker arms & fulcrums, mounted to the factory slotted pedestals, with adjusted geometry (fulcrum height), fastened with ARP #641-1500 bolts & #200-8587 washers. This is good for hydraulic tappet valve train up to about 0.550" net valve lift (rated by Ford to 0.615" valve lift).
(6B) All other valve train applications (especially those needing adjustability) should choose between Yella Terra YT6321 rockers or T&D Machine #7200 rockers. Both mount solidly to the cylinder head with 7/16" bolts and use push rod cup style adjusters.
(7A) Flat tappet cams: The camshaft should be ground on a best quality iron cam core, it should be ground with a guaranteed 0.002" lobe taper, it should receive the best quality hardening treatment (nitriding), and it should receive the best quality lobe polishing.
(7B) Roller cams: The camshaft should be ground on a core manufactured from a material which is compatible with standard OEM distributor gears or compatible with commercially available steel distributor gears. There are steel gears available which are compatible with the proper steel camshaft cores. Don't use bronze gears for a street engine. Therefore don't use cams requiring bronze gears in a street engine. Bottom line, the distributor gear must be compatible with the material the camshaft is made of.
(8) Raise the static compression ratio to achieve 7.6:1 to 7.7:1 dynamic compression (compatible with 91 octane pump gas).
(9) Breakerless ignition (20° centrifugal @ 3000 rpm, 16° to 18° initial, 10° vacuum adv. using ported vac.)
(10) Good oil pan (9 quart), windage tray, high volume oil pump pick-up, etc
(11) Achieve more than 50 psi hot oil pressure from 2000 rpm (60 psi is the target)
(12A) Limit oil to the valve train; tappet bore bushings w/0.060" orifices is always the preferred method.
(12B) In hydraulic tappet applications oil to the valve train can be limited via push rods with 0.040" restrictions.
(12C) In solid flat tappet applications oil to the valve train can be limited via special oil limiting tappets.
(12D) In solid roller tappet applications it is best to assume that tappet bore bushings are a necessity. Choose solid roller tappets that roll on solid bushings instead of needle bearings, and that employ forced lubrication instead of splash lubrication. Such as Isky #3972-RHEZ.
(13) Replace the crank damper (Romac #0203, ATI #918900, BHJ #FO-EB351C-7). The Romac damper is a nice damper but it is not bonded.
(14) If the flywheel requires replacement use a lightened steel flywheel. The Yella Terra YT9902 flywheel, weighing 26.4 pounds, is the lightest steel flywheel known to me. It is an external balance flywheel drilled for "long style" pressure plates.
(15) Replace the con-rod nuts (ARP #300-8371 nuts)
(16) Forged round skirt pistons (Ross #80556). Caution: Some pistons, like the Ross pistons, have increased wrist pin height, they shall raise the compression ratio. Such pistons are best used with D1AE head castings.
(17) Clevite MS-1010P main bearings, fully grooved (requires 2 sets), 0.0020" - 0.0025" clearance
(18) Clevite CB-927P rod bearings, 0.0025" - 0.0030" clearance
(19) Pull the small pin from the rear crankshaft seal groove of the #5 main bearing cap, seal the pin's hole with a dab of sealant, utilize a 2 piece neoprene seal in place of the OEM rope seal.
If machine work is performed, consider paying the extra expense for durability
(1A) Align bore if a straight crank cannot be turned by hand when cinched down into a fresh set of lubricated main bearings.
(1B) Align bore if the main bearing saddles have excessive taper or run-out.
(2) If the machinist insists on machining the decks, agree to only enough required to level the decks front to back and equalize their height bank to bank. This should not require more than 0.015". The deck height should never be machined less than 9.200". Might as well have the deck surfaces finished fine enough for MLS head gaskets while you're at it.
(3) Index the cylinders to the crankshaft (NOT the decks) during boring. This procedure will net you quite a few additional horsepower and better durability ... and the cylinders will be compatible with pistons having floating wrist pins.
(4) Choose full round skirt flat top pistons, such as the Ross #80556 pistons. The Ross pistons utilize modern thin piston rings and are drilled for wrist pin oiling. The full round skirt pistons exert their thrust forces over a wider area of the cylinder wall, they apply less stress on the "thin" cylinder walls of the production cylinder block. However, caution: The Ross pistons have increased wrist pin height, they raise the compression ratio, they are best used with D1AE head castings.
(5) Install 16 tappet bore bushings & 5 cam bearing restrictors
(6) Regrind the crank for 0.0020" - 0.0025" main clearance; 0.0025" - 0.0030" rod clearance. There should be no taper or run-out on the journals afterwards.
(7) Tufftride & micro-polish the crank
(8) Dynamically balance the reciprocating assembly
Performance (volumetric efficiency)
(1) Dual plane intake manifold (Blue Thunder or the older Shelby version recommended for 4V engines). Full height plenum, do not cut the plenum.
(2) Holley style carburetor, 735/750/780 cfm, center hung fuel bowls, vacuum secondaries, electric choke, street performance calibration, annular booster venturis.
(3) High lift camshaft (i.e. approx. 0.550" net lift) with 276°/286° advertised duration, 112°/117° lobe centers (or 282°/282° advertised duration, 109°/119° lobe centers), 114° LSA (112° minimum), 54° overlap (60° maximum).
(4) Camshaft should open the exhaust valve at 80° BBDC and close the intake valve at 70° ABDC (based on advertised duration).
(5) Headers, preferably tri-Y style, 1-7/8" to 2" diameter primaries for heads with 1.71" exhaust valves, 1-3/4" diameter primaries for all other heads.
(6) 2-1/4" to 2-1/2" tail pips, decent mufflers.
Thanks...LOTS of good advise.
Everyone should read this thread if you are considering any major engine work. It should be carved in stone on a tablet. BRAVO!
Very informative, anyone new to the Cleveland will benefit tremendously from your work George, well done, facilitating!!!!!!!
|72 Pantera #3165|
73 Maserati Bora
04 Ferrari 575M
94 Ferrari 348 Spyder
67 Mustang Fastback
67 Mustang Conv.
76 Cobra II
When I built my 351C I tried to do a lot of research. I found little as compared to my 351W project. I followed most all the advice stated in the above Sticky when it was under a different title. The result is a very powerful 351C that holds very good oil pressure. The bushings and restrictors were very easy to install.
George, Thank you for all your help with my build.
It might be here, if it is I've missed it...
Is there an EASY way to determine engine code and heads?
the cast numbers on my block are
2K18 (date ?)
stamped 01527 (I've lost the tag with engine number to verify)
on the front inside corner of the heads is a "4"
under the valve cover I only found
2J2 (date ?)
circle with L5316
large C with F in center
would the head cast number be under the intake?
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