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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. Cowboy Smiley


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


Part 1: Terminology and Production Information
Part 2: Durability
Part 3: State of Tune

Part 4: The Combustion Process (thermal efficiency)
Part 5: Induction System (volumetric efficiency)
Part 6: Exhaust System
Part 7: Camshaft and Valve Train

Part 8: Preparing A 351C Racing Engine


NEW 2022 Link: Making the 351C Super Stock
The 351C lubrication system
Link: Rebuild - Freshen Up - Overhaul
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


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



Outside of the R code engines ('71 351 Boss & '72 351 HO) all US manufactured 351 Cleveland short blocks were basically identical with three exceptions: (1) H code & M code engine blocks had two bolt main bearing caps, Q code engine blocks had 4 bolt main bearing caps; (2) the 1970 through 1972 pistons were flat top pistons with 3 cc dome volume, the 1973 and 1974 pistons were dished pistons with 8 cc dome volume; (3) the Q code engines had a bit larger/heavier crankshaft dampers. The nominal deck clearance for ALL H code, M code, and Q code engines was 0.035 inch.

The US 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 high rpm operation (>7500 rpm). This didn't stop people from using the crankshaft for racing however! The cranks would be internally balanced with "mallory metal" and then 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.

We've established the cranks are ALL tough iron alloy castings. They are externally balanced and should be internally balanced for racing. They are partially counter-weighted which means they should be restricted to about 7200 rpm usage. There are two other considerations. The first is the oil passages drilled within the crankshaft. The oil passages are suspected to contribute to the Cleveland's lubrication woes. There is no other way to explain why the connecting rod bearings for cylinders 2 and 7 always fail the first or the worst. And finally, cast iron has dampening properties which steel lacks. Consequently the cast iron crankshaft is "gentler" on the cylinder block than a steel crank. The factory iron crankshaft maximizes the durability of the thin wall cylinder block.

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 R code engines were prepared this way by the factory.

Ford testing in the 1960’s found a rod length to stroke length ratio of 1.72:1 provides the ideal rod length for an engine operating up to about 7500 rpm. It reduces cylinder wall thrust as much as possible while being just short enough to avoid the onset of induction system lag. The 351W has a rod length to stroke length ratio of 1.70:1. However, the 351C has a rod length to stroke length ratio of only 1.65:1. The 351C connecting rod (5.780 inches length) was intentionally shortened a bit to improve "street" (low rpm) performance. 

Connecting rod length also impacts the distance the piston is pulled out of the bore at BDC. The piston of a stock 351C is pulled so far out of the bore at BDC that the wrist pin is partially exposed (0.126 inch) beyond the lower edge of the bore. This is not ideal, but has never been an issue with street engines. 6.00 inch long Chevy connecting rods were once considered de rigueur when preparing a 351C for racing however. The rod length to stroke length ratio with Chevy rods was 1.71:1 and the wrist pin remained 0.087 inch above the lower edge of the bore at BDC.

It is best if the piston is never pulled out of the bore so far as to expose the wrist pin because the axis of piston rocking is the wrist pin axis. The further the piston is pulled out of the bore, the less surface the piston skirt has to press against. The piston skirt can’t perform as intended if it has no cylinder wall to press against. Thus the further the piston is pulled out of the bore at BDC the more it shall rock within the bore. Piston stress and wear will increase. At some point operating friction begins to increase. At a further point visible piston damage will begin to occur. By that point connecting rods can begin to bend or twist as well.

The R code engines ('71 351 Boss & '72 351 HO) were also equipped with standard production castings & forgings which received special inspection. The crankshafts were brinnel tested to insure a certain degree of hardness, and the connecting rods were magna-fluxed to insure there were no flaws in the forgings.

The 1971 Boss blocks were equipped with four bolt main bearing caps but were otherwise no different than the other production blocks. Four bolt main bearing caps became a standard production option the following year (Q code & R code engines). The connecting rods were shot-peened and equipped with 180,000 psi chromoly rod bolts;

The R code engines were equipped with lighter - stronger forged pistons with rounded skirts whereas all other 351 Cleveland short blocks were equipped with cast pistons. The H.O. engines also had crankshaft dampers which were bonded and heavy enough for engine speeds above 6000 rpm whereas all other 351 Cleveland engines had unbonded, lightweight or medium-weight crankshaft dampers.

Since ALL of 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; tappet bore bushings and fully grooved main bearings 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 7200 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 7200 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.

  • Main bearing lubrication does not have priority or first claim to the oil discharged from the oil pump
  • An excessive amount of oil flowing to waste in the tappet clearances reduces the amount of oil available for lubricating the crankshaft
  • An excessive amount of oil flowing to the valve train or the camshaft bearings also reduces the amount of oil available for lubricating the crankshaft

(2) The large ports in the walls of the tappet bores create three additional problems.

  • The ports allows too much oil to flow to waste
  • The ports create tappet compatibility issues, possibly allowing too much oil to flow to the valve train
  • The ports allow cavitation resulting from the motion of the tappets to spread within the oil passages.

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.




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
62° overlap
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.


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.


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
58° overlap
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.


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.


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

Last edited by George P
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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.


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

Under Development

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.


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


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


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.


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.

Pantera Exhaust

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:


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.


If you're interested in one of the Ford camshafts, three of them are still available via the aftermarket.


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:

  • Opening the exhaust valve too late causes high rpm torque to fall-off like a brick if there is any exhaust system back-pressure (mufflers). You can see from the factory cam specs the engine likes the exhaust valve to open by at least 80° BBDC. The SVO cam, which opened the exhaust valve a little later (77° BBDC), had a power band that was known to flatten out early. This is one reason why I've made a note about installing the cam advanced.
  • Too much over-lap softens low rpm torque, makes the engine idle rougher, and decreases vacuum at idle. None of the factory cams have more than 62° overlap. The big intake valve and canted valve geometry of the 351C 4V increases the interaction between the exhaust port, the combustion chamber, piston motion, and the intake port during the overlap period. This is a good thing for a race engine, overlap can be used to scavenge exhaust gases and get the intake charge flowing early which improves volumetric efficiency. Drivability (i.e. performance below 3000 rpm) isn't important to a race engine. While amplifying the effects of overlap is a good thing for a race engine, its a bad thing for a street engine. Small increases in overlap quickly diminish the drivability of an engine equipped with 4V cylinder heads. It is not the size of the intake port that hurts low rpm performance and drivability as people often assume, its the size of the intake valve and the manner in which it amplifies the effects of overlap. In designing or selecting a 351C 4V street camshaft its important to place emphasis on minimizing overlap in order to optimize low rpm power and maintain drivability.
  • Narrow lobe centerlines (LSA) create the first two conditions. Combined they have the effect of making the torque curve (power band) narrower and steeper. Since the effects of overlap are amplified, the performance of a 351C 4V when equipped with cams having 112° to 114° LSA is on par with the performance of other engines equipped with cams having 108° to 110° LSA. This is an aspect people can't wrap their head around. The indoctrination enthusiasts get via magazines, internet and television makes them think they are losing out on something if they install a camshaft having wide lobe centerlines (LSA). This just isn't so with a 351C equipped with 4V heads, it makes an abundance of hard hitting horsepower even with a camshaft having 114° lobe centerlines. The big 2.19" intake valves change the rules.
  • Closing the intake valve too late causes low rpm reversion while lowering dynamic compression. Modern cam lobes have lobe intensity specs that are more aggressive (smaller numbers) than the factory cams, so closing the intake valve by 70° ABDC seems to be a reasonable limit. The Cobra Jet cam and the D1ZZ-BX cam both closed the intake valve at about 40° ABDC based on duration at 0.050", that's another calculation I make to check cams in these days of faster cam lobes.

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.


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.


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
Last edited by George P


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.


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

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 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 five conditions:

(1) The cam lobes are ground with insufficient taper, therefore the tappets do not rotate as they are lifted & lowered by the cam lobes. I am told this is why they fail, and I believe it is intentional. Most of the problems seem to centered around one cam company.

Other possible problems:

(2) A lubricant issue (i.e. insufficient wear protection)
(3) A quality control issue (i.e. the parts were not made of the same materials used decades ago or were not surface hardened properly)
(4) 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)
(5) 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 5 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.


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.


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 ************, 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 440 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 to about 10.5:1 (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 10.5:1 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. Some crankshafts also offer improved rod bearing lubrication passages, which is a 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). The motor should be set up for 10W or 15W oil and use fully grooved copper-lead alloy main bearings (Mahle/Clevite MS-1010P). Acquiring fully grooved bearings can be accomplished using the upper halves of two sets of standard bearings. 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 ************ 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!


Last edited by George P

Rebuild - Freshen Up - Overhaul
Assuming it is in sound working order, the limits of a 351C with factory machine tolerances, factory cast pistons, factory connecting rods, factory balanced reciprocating assembly, and hydraulic tappet valve train are 425 horsepower, 6500 rpm, and 8.0:1 dynamic compression or 10.5:1 static compression (whichever comes first). At that point the cast pistons, the engine's thin cylinder walls, the connecting rod fasteners, and the factory valve train are highly stressed. The factory fuel pump and the factory clutch are also pushed to their limits.

425 naturally aspirated horsepower is on par with the top pre-emissions (pre-1968) production street engines of the muscle car era: Ford's 8-bbl 427 (R code), Chrysler's 8-bbl 426 Hemi, Chevy's 4-bbl 396 (L78), and Chevy's 4-bbl 427 (L72) were all rated 425 horsepower. Pontiac's 6-bbl 421 Super Duty was the lowest rated at merely 376 horsepower, and Chevy's 6 bbl 427 (L71) was the highest rated at 435 horsepower.

The folks at De Tomaso built Pantera race engines using production castings. These were race prepped engines with forged pistons, chromoly rods, internally balanced cranks, and solid tappet cams. To survive a season of European endurance racing the engines were limited to 440 horsepower, 7000 rpm, and 10.5:1 static compression. Their strategy was not to make as much horsepower as possible but to stay within the engine’s limitations for that application (European endurance racing). In ten years of racing, 1974 to 1984, Odoardo Govoni won 3 Italian championships and 1 European championship piloting the group 4 Pantera chassis # 2873.

To push an engine to its limits, or to maximize an engine's limits, it should be in sound working order. Its bearings, rings, and valves should be in good condition, its reciprocating assembly should be factory balanced or better, and it should be in a good state of tune. Thus a rebuild, overhaul, or "freshen-up" is usually in order. Typical of a rebuild, if pistons are needed they would be low cost OEM style cast replacements.  Machine shop services would be relied upon only for cylinder head servicing and connecting rod servicing. Crankshaft journals and cylinder bores would be used "as-is" when possible; bearing journal reconditioning and cylinder boring services would only be enlisted when necessary.

An engine that is "merely" rebuilt or freshened-up shall not be as durable as it could be if it were race prepped or blue printed. Nothing is done to extend an engine's limits. There is nothing wrong with that. But, as a consequence, the as-built limitations suggested above remain the boundaries to work within.

Ten Emendations to Include in a Rebuild
The 351C has 3 recurrent issues which can be quite destructive:
(1) The heads of the factory valves break-off their valve stems from time to time and bang around inside the cylinders.
(2) The connecting rod bearings are under-lubricated which damages the bearings to the extent that hot oil pressure is impaired.
(3) The threads strip out of the connecting rod nuts, the big end caps come loose, and the connecting rods bang around within the cylinders.

Taking steps to prevent those issues from happening is logical, especially since two of the issues simply require the replacement of parts. The third requires the installation of bushings in the engine block and increasing bearing clearances. The steps for preventing all 3 issues should be taken when an opportunity (like a rebuild) arises to make them.

(4) An upgrade to a breakerless electronic ignition can be accomplished with Ford parts, with the exception of the wiring harness. Immediately noticeable upon installation of a breakerless ignition, the engine starts quicker, runs smoother, and has better low rpm power. Power actually improves at all engine speeds. The ignition is a significant improvement over any breaker point ignition, including a dual point ignition. There are no points to adjust, thus maintenance is reduced.

Those four improvements are appropriate for any and all 351C rebuilds. The remaining  six emendations pertain to performance engines.

(5) An oil pump cannot provide an uninterrupted supply of oil to the engine unless its suction is constantly submerged in a bath of non-aerated oil. An inadequately designed oil pan can exacerbate the engine's lubrication problems mentioned above (issue #2). A high-capacity oil pan with baffles, hinged doors, a windage tray, and a scraper enables the lubrication system to better cope with the high G-Force capabilities (braking, cornering, acceleration) of the Pantera's chassis.  Such an oil pan should have been standard equipment for the Pantera. De Tomaso equipped the GTS, GT5, and GT5-S Panteras with high-capacity oil pans and windage trays.

(6) A prerequisite to raising the compression ratio is a carburetor with a richer A/F ratio. Using an emissions-tuned carburetor with raised compression would result in knocking, pinging, and detonation. There are no factory carburetors with the proper calibration.

Besides being a prerequisite to raising compression there is also horsepower to be gained by replacing the 4300 series carburetors. The 1970 carburetors were down 3% (down 10 HP out of 310 potential HP), the 1971 carburetors were down 8% (down 25 HP out of 310 potential HP), the 1972 and 1973 carburetors were down 20% (down 70 HP out of 350 potential HP), and the 1974 carburetors were down 17% (down 61 HP out of 350 potential HP). De Tomaso equipped the GTS, GT5, and GT5-S Panteras with 650 cfm (non-emissions) Holley list R4777 carburetors.

Engines equipped with factory spread bore carburetors will need to replace the "spread bore" intake manifolds with dual plane manifolds designed for "square bore" carburetors.

(7) The 351C was a performance engine manufactured in an era of declining performance, and declining compression ratios. Improving (restoring) the 351C's performance has been a common theme since the engine was new. Raising the compression ratio improves the engine's dynamic performance (throttle response, acceleration, low rpm torque). This is the most important change to make for improved performance. Zero to sixty times improve by 2 to 3 seconds.

(8) The factory push rods were flimsy. Ford and other manufacturers designed them that way intentionally as a sort of mechanical rev limiter. To gain rpm at the upper end of the power band, and to extend the rev limit, stiffer push rods are needed.

(9) The connecting rod fasteners are the most stressed and most critical fasteners in the engine. If 350 horsepower was expected to be adequate the OEM fasteners would also be adequate. It is prudent however to improve the fasteners in preparation for increased output and increased rpm, even Ford installed chromoly connecting rod fasteners in the 351 Boss.

(10) Ford also installed a heavier, fully bonded damper on the 351 Boss. Since the damper is most likely overdue for replacement, it seems logical to purchase a damper similar to the 351 Boss damper for this or any "performance" application. The 351 Boss damper has not been available for decades, so we'll have to resort to an aftermarket damper of similar or better design. A good one will be suggested.

Factual Compression Ratio
There are two things to become accustomed to as you continue reading.

First, no auto manufacturer's sales department ever published dynamic compression ratios, although their engineering departments used dynamic compression ratio to determine what fuel octane an engine required.

Second, the static compression ratios being quoted are factual, real-world compression ratios; not the grossly exaggerated advertised compression ratios that US Ford published through 1971.

Most of the high compression engines of the 1960s operated on "premium" gasoline, that indicates their dynamic compression ratios were in the range of 7.0:1 to 8.0:1. Fuel rated "regular" was for engines having 7.0:1 dynamic compression or less. Fuel rated "super-premium" was for engines having dynamic compression greater than 8.0:1. Very few engines were ever designed for "super-premium" fuel. Ford specified premium fuel for the 1970 M code, 351-4V. The sales department's claim that it had 11:1 compression was never possible.

Mass produced engines are built with tolerances. The 1970 M code engine had between 9.30:1 and 10.02:1 static compression, 9.65:1 being the nominal value. The engine's dynamic compression was between 7.14:1 and 7.67:1, 7.40:1 being the nominal value.

How Much to Raise the Compression
Raising the compression of an engine improves its "dynamic performance" in terms of low-end torque, pep, acceleration and throttle response. As a consequence, having a strong performing engine is secondarily about how much horsepower the engine makes … it is primarily about how much compression it has.

Raising the compression of an engine improves performance so well that many enthusiasts assume it must significantly increase horsepower …  but it doesn't. For example, assuming an engine produces 350 horsepower:
• Raising the compression from 8.0:1 to 10.0:1 increases output by 11.0 HP.
• Raising the compression from 8.5:1 to 10.0:1 increases output by 7.4 HP.

Raising compression is also a condition of diminishing returns:
• Raising the compression from 8.0:1 to 9.0:1 increases output by 6.6 HP.
• Raising the compression from 9.0:1 to 10.0:1 increases output by 4.5 HP.
• Raising the compression from 10.0:1 to 11.0:1 increases output by 3.1 HP.
• Raising the compression from 11.0:1 to 12.0:1 increases output by 2.3 HP.

There are limits to how much the compression can be raised. The 351C was originally designed to operate on premium fuel. That is 91 octane US/Canadian fuel (R+M/2) or 95 octane international fuel (RON). Thus, it was intended to have dynamic compression no greater than 8.0:1. The two 351C factory engines with the highest compression, the 1970 M code and the 1971 R code (351 Boss), had in fact 7.40:1 dynamic compression and 7.43:1 dynamic compression, respectively (nominal values).

The Cleveland is infamous for having cylinder walls designed as thin as possible; the engineers didn't leave any "wiggle room". The Cleveland cylinder wall thickness specification was 0.160 inch ± 0.030 inch. The cylinder walls could have been as thin as 0.130 inch in spots, right off the assembly line. That is literally half as thick as they should have been. This limits how high the compression ratio can be safely raised without incurring cylinder wall damage. That limit is traditionally 10.5:1 static compression.

In regard to compression, the Cleveland's upper limit is best stated as 8.0:1 dynamic compression (based on the fuel) or 10.5:1 static compression (based on the cylinder walls), whichever comes first.  Those limits, as they impact street operation on premium fuel, employing short duration street cams, aren't bad at all, they pose no handicap to performance.

With high performance (Cobra Jet) street engines, short intake duration and early (advanced) cam timing combine to close the intake valves early, commonly between 66° and 72° ABDC. A static compression of 10:1 produces dynamic compression ratios between 7.92:1 and 7.53:1, respectively. However, 10:1 compression is not exactly 10:1 compression. It varies based on the diameter (over-bore) of the cylinders. 10:1 static compression can actually be as low as 9.80:1. A "worst-case" scenario combining 9.80:1 static compression with a 72° IVC produces a dynamic compression ratio of 7.40:1. That's still OK.

How Best to Raise the Compression
To raise the compression of a 351C ALL Cleveland cylinder heads (excepting the Australian 302C heads) need some amount of milling to reduce their chamber volumes.

Raising the engine's compression with open chamber heads can be accomplished in two ways. The first way requires zero-decking the block (lower the decks by 0.015-inch, raise the compression height by 0.020-inch). The second way employs pop-up dome pistons. Either method requires an expensive set of pistons and re-balancing the reciprocating assembly.

BUT raising compression with quench chamber heads doesn't require milling the block, an expensive set of pistons, or re-balancing the reciprocating assembly. If dished pistons need to be replaced, an inexpensive set of cast flat tops which weigh the same as the dished pistons (about 616 grams) is all that is needed.  The simplicity and lower expense of raising compression with quench chamber heads  makes them the best method for raising compression during a simple rebuild.

Ford produced the following quench chamber heads.
US Ford 1970 M code, D0AE castings, 4V ports, 62.8 cc chambers
US Ford, 1971 M code, D1AE castings, 4V ports, 66.1 cc chambers
Australian Ford '72–'75 Y code (302C), 2V ports, 57.9 cc chambers
Australian Ford '76–'84 P code (302C), 2V ports, 57.9 cc chambers?
• US Ford Motorsport 1986, p.n. M-6049-C351, 2V ports, 62.0 cc chambers

To make 10:1 compression with any of the US  manufactured quench chamber cylinder heads start by milling the heads to achieve 61 cc chamber volume. Milling quench chamber heads 0.006-inch reduces chamber volume by 1 cc. D0AE heads require nominally 0.010-inch milling, D1AE heads require nominally 0.030-inch milling. Employ production cast flat top pistons, and employ 0.038-inch head gaskets. If over-bore pistons with full 1.650 inch compression height and only 3 cc dome volume are used then 0.047-inch head gaskets would be necessary. This is unlikely if off-the-shelf cast pistons are employed.

The Australian 302C heads have smaller combustion chambers. To make 10:1 compression they do not require milling, use them as is (58 cc chamber volume). Employ production cast dish top pistons, dish top pistons have 5 cc more dome volume than flat top pistons. Employ 0.038-inch head gaskets. If the heads are milled then 0.047-inch head gaskets would be necessary.

The following 351 Cleveland specific rebuild information is presented in the format of a "budget conscious" home mechanic style rebuild.

Cylinder Heads
It is assumed that 4V quench chamber (M code) heads shall be utilized. If they are not on hand, acquire a set.

Recurring Problem # 1
The 351C has an issue with the valve heads of the factory valves breaking-off their valve stems from time to time, they bang around inside the cylinder and wreak havoc on the engine. To prevent this from happening, and for peace of mind, t
he factory valves must be replaced.

The factory valves were two piece valves; the stems were welded to the valve heads. Most valves are manufactured that way. There are people who have been led to believe that in order to be reliable replacement valves "must" be of one piece construction. However, the two piece construction of the factory valves had nothing to do with their failure. When the valve heads crack and break off the stems, this happens below where they are welded together. The problem with the Ford valves was that the valve heads were brittle.

D0AE, D1AE and D1ZE 4V heads are equipped with the large 2.19/1.71 valves. All other Cleveland heads, including the D3ZE 4V heads, are equipped with 2.041/1.655 valves.

Replacement valves can be whatever type or brand are preferred, so long as they are not another set of factory valves! A set of steel valves from Melling or Sealed Power would be satisfactory. For a performance oriented engine stainless steel valves make a good choice because they are lighter weight, swirl polished, and have single groove style stems. Stainless steel valves (and titanium valves) are made of softer materials, the valves tips should be equipped with steel inserts. The valves sold by Manley Performance are excellent suggestions … but there are lower priced alternatives.

The first machine shop task is servicing the M code heads. The heads may be needed once servicing the block is underway.

Machine Shop Task #1:
Servicing the heads is a five-part task: (A) hot tank, (B) check for leaks, (C) mill the heads, (D) install bronze valve guides, and (E) service the valve seats.

The heads should be milled to achieve 61 cc chamber volume; mill D0AE castings nominally 0.010 inch, mill D1AE castings nominally 0.030 inch.

Bronze Guides & Positive Seals

The bronze valve guides should be the type which are flanged at the top for "positive seals". The old cast iron guides sticking up above the spring pedestals are eliminated. The cast-in guides inside the valve pockets can be eliminated too, that is a matter of personal preference. Ford's endurance racing specs for valve stem to valve guide clearances are:
Intake valves: 0.0007 inch to 0.0018 inch (0.0012-inch nominal)
Exhaust valves: 0.0011 inch to 0.0022 inch (0.0016-inch nominal)

Reaming all the valve guides for 0.0014-inch clearance is one strategy. The production specs were nominally 0.0018-inch for the intake valves and 0.0023-inch for the exhaust valves.

If the heads require new valve seats then stainless steel valves (and titanium valves), which are made of softer materials, should be complemented with iron or beryllium copper ($$$) valve seat inserts. Hardened steel or stellite inserts should be used with steel valves.

3 angle valve seats

The Cleveland ports are well designed and perform competently "as-is". With a little "clean-up" and a "proper" three angle valve job air flow can be improved by about 15%. But, the port's performance cannot be improved by modification, only undermined. Thus as little material as possible should be removed.

Three angle valve seats should appear as three thin - concentric - rings, such as the cuts in the close-up picture. Take note that the valve pocket in the center picture has been hand smoothed, not cut with a machine. Cutting was restricted to the valve seats. Very little material has been removed.

The valve pocket has very specific dimensions and an intentional shape (profile). There is a venturi (throat) above the valve seats. The valve pocket has "some" cutting above the valve seat performed by the factory. That cutting roughly defines the pocket's throat, but it can use some improvement. The 4V intake valve pocket's throat (1.74-inch diameter) should not be enlarged or eliminated. The valve pocket cannot be serviced or improved with a valve seat cutting machine. Any wide cut, tapered like a funnel or straight walled like a barrel, RUINS the valve pocket and undermines port performance. The head in the left hand picture above appears to have been additionally cut, beyond the factory machining. If so, then it has been butchered. That head would honestly perform better if the valves had just been lapped-in.

Choose a machinist who will restrict his valve seat cutting to three thin - concentric - rings.

The original factory machining left ridges and sharp edges around the valve seats in the combustion chamber and in the valve pocket. For the best air flow a little bit of preliminary hand blending is helpful to knock down and blend the ridges left by machining, and a little emery cloth work is sufficient to round-off the sharp edge in the throat. The backside of the valve pocket, opposite the short turn radius, can use a little additional blending above the throat, to better define the venturi.

The valve seat angle should be cut at 45°.

The valve seat should be blended into the combustion chambers with a 30° cut. The peripheral size of the 30° cut is limited by the adjacent valve. The inner diameter of the 30° cut sets the peripheral size of the valve seat, which should be near the outer edge of the valve. This cut must negotiate those limits, thus it is necessarily narrow, but should be about 1/16-inch wide.

The valve seat should be blended into the valve pocket with a 60° or 70° cut. Ford literature recommended both. 60° may be a little more resistant to human error. The width of the 60°/70° cut should be the same around the entire circumference, about 1/8 inch. The periphery of this cut sets the width of the valve seat. To improve air flow without undermining durability the width of the intake seat should be 0.060-inch to 0.070-inch, or about 1/16 inch. Much narrower than in the left hand picture above. To provide adequate heat transfer the width of the exhaust seat should be 0.080-inch to 0.090-inch, or a bit more than 5/64-inch but a bit less than 3/32-inch.

Finally, valve seat run-out should be 0.0010 inch or less. The run-out spec is important for durability.

Engine Block
Degrease the block. Don't neglect the cam bearing oil passages. They always get plugged with sludge.

relief 2

The M code heads were equipped with 2.19-inch intake valves. If the engine's   block doesn't have reliefs ground into the tops of the cylinders for 2.19" intake valve clearance, those reliefs will have to be ground. This is when the serviced heads are needed. If the cylinders are going to be bored, this should be accomplished before boring them.

Assemble one head using valve train checking springs. Use that head on both banks to establish where to grind the 2.19-inch valve reliefs, and to check for clearance afterwards. Snugly bolt the head to the engine block with dowels in place and flip the block upside down to mark where to grind.

> A Decision must be made whether or not the cylinders should be bored. The degree of scoring is one factor to consider, cylinder bore taper is another. The factory wear limits are:
Maximum cylinder bore out of round: 0.005-inch
Maximum cylinder bore taper: 0.010-inch
Piston to cylinder bore clearance: 0.0014-inch to 0.0022-inch

That much taper is unlikely; but it needs to be measured to be sure. Severely tapered cylinders would need to be bored and honed for oversize pistons. If the cylinders must be bored DO NOT jump immediately to a 4.030-inch over-bore. Over-bores should be kept to a minimum due to the thin cylinder walls; Clevelands can usually be cleaned-up with merely 4.010-inch over-bores.

If the cylinders are not being bored, then ream the ridges at the top of the cylinders and hone or deglaze the cylinder walls. Clean the cylinder walls well afterwards.

If the engine has dish top pistons, or if the existing pistons are badly scored, substitute them with a new set of factory style cast flat top pistons (1.650-inch compression height). The factory pistons have steel struts in the skirts and are a bit heavy. The replacement cast pistons should be designed the same and should weigh the same as the OEM pistons (about 616 grams), thus the crank won't need rebalancing. Seek out pistons manufactured in North America (Silv-O-Lite, Sealed Power, Clevite, etc.).

Replacement pistons for standard 4.000-inch bores should have 1.650 inch compression height and 3 cc dome volume (valve reliefs). For over-size bores you should expect them to have less compression height OR increased dome volume (slightly dished domes), but not both. These are the piston manufacturers'  methods to prevent an increase in compression resulting from an over-bore. The following piston dimensions maintain the compression ratio for various bores:

• Std. 4.000 bore: 1.650-inch compression height and 3.0 cc dome volume
• 4.010 over-bore: 1.648-inch compression height or 3.4 cc dome volume
• 4.020 over-bore: 1.646-inch compression height or 3.8 cc dome volume
• 4.030 over-bore: 1.644-inch compression height or 4.2 cc dome volume
• 4.040 over-bore: 1.642-inch compression height or 4.6 cc dome volume

Machine Shop Task #2:
Replacing the pistons will require the help of a machine shop to remove and re-install the pressed pins. The spec for the minimum amount of force to remove the pressed pins is 1800 lbs. If the pins are removed with less force than that there is a problem. Pins are normally installed by preheating the small end of the connecting rod.

If the pistons aren't being replaced, clean the ring grooves of the existing pistons. Those pistons need to go back into the bores they came out of, clocked the same way they came out.

If the cam bearings are being replaced, installation of the no. 1 cam bearing requires special attention. If the bearing is not installed properly then the camshaft thrust plate, the timing gear, and the fuel pump eccentric will be unlubricated. The slot in the #1 cam bearing must be oriented to connect the cam bearing oil supply passage and the thrust plate oil passage. Once the slot in the bearing is aligned to connect the two oil passages the bearing must be installed to a depth of 0.003 inch to 0.005 inch within its bore, beyond flush with the front machined surface, as measured with a straight edge and feeler gages.

This is a good time to point out that the slots or holes in the camshaft timing sprocket allow spray lubrication from the thrust plate oil passage, located behind the cam sprocket, to reach the fuel pump eccentric on the front of the cam sprocket. Some camshaft timing sprockets lack holes or slots, don't purchase one like that.

Recurring Problem # 2
The 351C has an issue with under-lubrication of the connecting rod bearings, and low hot oil pressure (should be 50 to 70 psi at 2000 rpm). As the bearing wear increases, the hot oil pressure decreases … and so do the engine's durability limits. The engine's limits (425 horsepower, 6500 rpm) depend upon the engine's bearings being in good condition.  This issue shall be dealt with in this section and the next three sections.

tappet bore bushings

Install tappet bore bushings in all 16 tappet bores. Suggested orifice size is 1/16th inch (0.062-inch). This can be done at home, the kit is $400. Clean the block again, after enlarging the tappet bores, but before installing the bushings … and then clean it a third time after the bushings are reamed to size. It must be assumed every particle of metal left behind will become a problem.

DO NOT install a restrictor in the oil passage supplying the left-hand bank of tappets. The single restrictor has led to the random collapse of hydraulic tappets on the left-hand side of the engine. It's only needed when merely 8 tappet bore bushings are installed (in the right hand tappet bores), and it is only viable with solid tappets. It's not needed at all when bushings are installed in all 16 tappet bores.

The benefits of installing 16 tappet bore bushings include:
Hot oil pressure at 2000 rpm shall be within spec (50 to 70 psi).
Minimizes the amount of oil loss due to leakage at all 16 tappet bores.
Maximizes the amount of oil supplied to the main bearings.
Prioritizes the lubrication of the main bearings.
Isolates the oil passages from tappet motion (cavitation).
• Prevents the random collapse of left-hand hydraulic tappets.
Resolves tappet compatibility issues on both sides of the engine.
Prevents an excessive amount of oil flowing to any part of the valve train.
• Equalizes the amount of oil metered to all 16 rocker arm/valve/valve spring assemblies.

"Carefully" and "lightly" chamfer the oil holes in the crank, clean the passages afterwards.

Crank Polishing

Polishing the journals of a nodular iron crank in the wrong direction shall raise microscopic burrs which make the journals abrasive, leading to bearing damage. So don't polish the crank unless which direction to polish the crank is understood. The correct direction will cause the burrs to lay-down as the crankshaft rotates.

> A Decision must be made whether or not the crankshaft journals should be reconditioned. The degree of scoring is one factor to consider. Here are the factory specs:
Main bearing journal OD = 2.7484-inch to 2.7492-inch (2.7488-inch nominal)
Main bearing journal maximum run-out = 0.004-inch
Main bearing journal maximum out-of-round = 0.0004-inch
Main bearing journal maximum taper = 0.0003-inch per inch
Rod bearing journal OD = 2.3103-inch to 2.3111-inch (2.3107-inch nominal)
Rod bearing journal maximum out-of-round = 0.0004- inch
Rod bearing journal maximum taper = 0.0004-inch per inch

New bearings should be heavy duty bearings (Clevite tri-metal or similar). Heavy duty rod bearings are not damaged as badly by under-lubrication as standard babbitt style bearings are. They will also stand-up better to the less than perfect journal surfaces if the crankshaft is not reconditioned.

The main bearings should be fully grooved. Fully grooved main bearings were once standard equipment in heavy duty bearing sets; they more than double the amount of oil supplied to the connecting rod bearings. The 351C has NEVER suffered any problems stemming from the use of fully grooved main bearings, only benefits. Considering the Cleveland's problems with under-lubrication of the rod bearings, fully grooved main bearings should be considered essential.

If the crank is being reconditioned by a machinist, specify 0.0025-inch to 0.0030-inch rod & main clearances.  If it is not being reconditioned then achieve those clearances by using oversize bearings. Those clearance specs are Ford's endurance racing specs. Ford's production specifications were too tight:

    Factory Main Bearing Clearance Spec
      • Desirable spec = 0.0010 inch to 0.0015 inch
      • Allowable spec = 0.0011 inch to 0.0028 inch
    Factory Connecting Rod Bearing Clearance Spec
      • Desirable spec = 0.0010 inch to 0.0015 inch
      • Allowable spec = 0.0011 inch to 0.0026 inch

If fully grooved mains can't be found in oversized kits, use the grooved upper shells from two standard (1/2 groove) oversized bearing kits.

Install the main bearings, the crankshaft, and the main bearing caps. Torque the main bearing cap inner bolts (1/2-inch) to 95-105 ft./lbs. Torque the main bearing cap outer bolts (3/8-inch) to 35-45 ft./lbs (IF the caps have outer bolts).

Make sure the crank can be rotated by hand in the bearings. Then measure the crankshaft’s end play. If the end play exceeds 0.010-inch the excess end play can be corrected using a no. 3 main bearing with thicker thrust faces, known as a “cranksaver” bearing.

Checking the flywheel is another task to perform while the crankshaft can be rotated by hand. If the flywheel’s assembled clutch face run-out is 0.010 inch or less, if it has the typical grooves and ridges but can be cleaned up by removal of no more than 0.045-inch material, then a simple resurfacing may suffice. However, if the flywheel shows signs of stress cracking or heat checking, if it would require the removal of more than 0.045-inch material to clean it up, if it is warped or has excessive clutch face run-out, then the flywheel should be replaced. The flywheel bolts are torqued to 75-85 ft./lbs.

Once all of that is checked and any problems are resolved, disassemble the main bearing caps, remove the crank, and install the rear main seal. If you're replacing the rope style rear crank seal with a split rubber seal, pull the little pointy pin from the seal groove in the #5 bearing cap. Put a dab of sealant in the hole the pin left behind.

The main bearings, the crankshaft, and the main bearing caps can be reassembled for the final time.

Motor Oil
“Low viscosity” motor oil promotes the flow of oil through oil passages, around the corners of intersecting passages, and through bearing clearances. There are other considerations however:
There is concern that if the oil viscosity is too low the grooves in the lower bearing shells of "fully grooved" main bearings may act like the rain grooves of tire treads and channel oil away from the bearing clearances.
The amount of oil flowing into the rod bearings increases as viscosity decreases, but the persistence of the oil wedge improves as viscosity increases.
1970 mass-production machine tolerances (concentricity, roundness, taper) are not good enough for ultra-low viscosity motor oil.

Thus, there is a balancing act involved with the selection of oil viscosity; low viscosity motor oils are advantageous, so long as the viscosity is not too low. To support the goal of maximizing the lubrication of the rod bearings low viscosity motor oil rated 10W30, 10W40, or 15W40 should be used. But, to prevent any of the expressed problems from arising oil viscosity lower than 10W30 is not recommended.

Valvoline VR1 10W30 synthetic oil is an excellent choice for an older engine like the 351C, especially one equipped with a flat tappet camshaft.

Connecting Rods
Install the piston and rod assemblies, less rings, in the appropriate bores and fasten them with new bearings to the rod journals. Its important to maintain the rod pairings for each connecting rod journal. Rods sharing the same journal are 1&5, 2&6, 3&7, and 4&8. Torque the rod nuts to 45 ft./lbs.

Cylinder Numbering

Connecting rod "side-clearance" is the source of "oil splash". Oil splash lubricates the thrust walls of cylinder bores, it also lubricates cam lobes and tappets. Most relevant to the subject of 351C lubrication, tight side-clearances will inhibit the flow of oil through the rod bearings.

The side-clearances between paired rods should be 0.018-inch to 0.022-inch as per Ford's endurance racing spec. The production spec was 0.010-inch to 0.020-inch with a wear limit of 0.023-inch. Measure the rod side-clearances, see how close they are to 0.018-inch before you decide how to proceed. I'd reluctantly let 0.016-inch ride, but no less.

One way to measure the side-clearances is with a dial indicator, tapping the rods lightly to be certain they are fully to one side and then the other. Another way is to spread the rods apart, then measure the gap with a feeler gauge. If the side-clearance is insufficient the inside mating faces of the rods (not the fillet sides) must be cut with a surface grinder. Half of the desired increase in side-clearance should be taken off each mating face of the pair.

Machine Shop Task #3:
Take the rods to a machinist in pairs (rods from the same journal), and have metal removed from the mating faces of each pair based on the side clearance measurements.

Recurring Problem # 3
While on the subject of the connecting rods, the 351C has an issue with threads stripped out of the connecting rod nuts.  The stripped-out nut allows the big end cap to come loose, the connecting rod bangs around within the cylinder, and severe engine damage occurs. To prevent this from happening the connecting rod nuts should be replaced with the ARP # 300-8381 nuts.

However, if the rods are being taken to a machinist, consider going one step further and install heavy duty rod bolts, then have the machinist resize the big ends while he has the rods. The additional expense is minimal. Whereas the factory rod bolts have a tensile strength of 150,000 psi and are rated for 45 ft./lbs torque, chrome moly rod bolts have a tensile strength of 180,000 psi and are rated for 50 ft./lbs torque.

Note: If the rod bolts are replaced the big-ends of the connecting rods MUST be resized.

ARP sells two chrome moly rod bolt kits: kit # 154-6003 (knurled shanks) and kit # 154-6403 (wave-loc shanks). For the ultimate in strength and clamping force another ARP “wave-loc” rod bolt kit, # 254-6403, is manufactured using their proprietary ARP2000 metal. Those rod bolts have a tensile strength of 220,000 psi and are rated for 55 ft./lbs torque. That one is a bit more expensive.

Machine Shop Task #4:
If the rod bolts are replaced have a machinist resize the connecting rod big ends. The Connecting rod big-end ID spec is 2.4361-inch to 2.4369-inch. Ask the machinist to aim for the minimum spec.

Now that the pistons and rods are done going back and forth to the machinist hand fit a new set of rings to each bore, install them on the proper pistons, then install the piston/rod assemblies in their proper bores and fasten them to their journals. The amount of torque (45, 50, or 55 ft./lbs) varies depending upon which rod bolts are being used, as previously discussed .

Getting all the rings fitted, all the pistons back into the block, and all the rods fastened to the crank is a milestone; light can seen at the end of the tunnel.

Ignition - Distributor
The next step is to prepare the distributor.

Install a breakerless electronic ignition, to avoid aftermarket parts Ford's Duraspark I or Duraspark II ignitions are suggested. Set the ignition for 16° to 18° static advance, and 20° centrifugal advance spanning 1800 rpm. The curve should start at 1000 to 1200 rpm and be "all-in" by 2800 to 3000 rpm.

20° centrifugal advance requires a 10L sleeve & plate assembly, or a sleeve & plate assembly with an 0.410 inch slot width.

That timing protocol (36° to 38° total advance) applies only to factory iron heads and flat top or dished top pistons. Pop-up dome pistons require more total advance, "Yates" (aka high swirl) combustion chambers require less total advance.

Increasing static advance improves an engine's low rpm pep. However, increasing static advance also makes an engine more and more difficult to crank during hot starts. Limiting static advance to 16° to 18° prevents the cranking effort from exceeding the torque of the starter motor.

Connect the vacuum advance to "ported" vacuum. Capping-off the vacuum retard connection inhibits operation of the distributor's vacuum mechanism, so leave it open to atmosphere.

If a Duraspark distributor is already in use, the centrifugal advance mechanism in Ford distributors should be serviced about every 50,000 miles.

Distributor Gear

Distributor gears may require replacement due to wear or camshaft replacement. The OEM gear was iron, whereas cams ground on steel cores require steel distributor gears.

Checking the distributor's end play and drive gear placement is most important whenever the distributor or distributor gear are replaced.
Verify that distributor shaft end play is between 0.024-inch to 0.035-inch.
Verify that the distance from the bottom of the distributor housing mounting flange to the bottom of the distributor gear is 4.031-inch to 4.038-inch when the distributor shaft is pulled away from the housing.

There is also an empirical method for testing the positioning of the distributor gear on the distributor shaft. This check must be performed prior to installation of the camshaft.
Insert the distributor in the block.
With the distributor housing fully seated against the block, verify that the distributor gear can be lifted off the support within the block at least 0.005”.
Next pull the distributor gear downward and hold it firmly against the support within the block. Pull up on the distributor housing and verify you can lift it from the fully seated position at least 0.005”.

This procedure verifies the distributor gear is not being jammed downward against the support in the block and it is not being prevented from contacting the support in the block either.

If the position of the distributor gear must be moved, drill the new hole 90° offset from the existing hole. The holes are 0.125” (1/8”) ID. The distributor gear should fit the distributor shaft very tightly, and should require a press to install it or remove it.

Set the distributor aside for now.

Valve Train Inspection
Specialized tools and parts are required to perform these inspections. A set of 8 valve train checking springs, a pair of adjustable (measuring) push rods, a spare head gasket identical to the gaskets that shall be used for final assembly, a pair of spare tappets identical to the tappets being installed, a degree wheel, a dial indicator with a magnetic or clamping stand, and calipers capable of measuring at least 9 inches.

Rocker Arm Geometry
The goal in setting rocker arm geometry is to minimize valve guide wear. Although the rocker arm's tip travels in an arc, it's motion should be most linear in relation to the valve stem’s axis where it contacts the valve tip. Optimized geometry minimizes side thrust on the valve stem which in turn minimizes valve guide wear.

rocker geometry

Three conditions are indicative of optimized rocker arm geometry:
• An imaginary line drawn through the rocker arm's axis and across the valve tip is perpendicular to the valve stem’s axis when the valve is 50% open.
• The rocker arm's overall contact patch on the valve tip is thinnest when geometry is optimum.
• The rocker arm's point of contact on the valve tip is furthest from the fulcrum when the valve is 50% open.

Assuming the engine is equipped with factory rocker arms, the rocker arm geometry "should not" need to be checked if a factory cam is being utilized. If an aftermarket camshaft (higher lift) is being installed, then rocker arm geometry should at least be given a cursory inspection.

rocker tip wear pattern

The Cleveland rocker arms are rigidly mounted on the cylinder heads; rigidly mounted rocker arms are also described as pedestal mount, saddle mount, or bolt down rocker arms. Rocker arm geometry can be checked and adjusted with the heads on the work bench; geometry is independent of push rod length. Rocker arm geometry is adjusted by raising or lowering the fulcrum height (aka stand, pedestal, or saddle height).

The heads will need to be fitted with valve train checking springs.

Factory Rocker Arm Geometry

Theoretically, the fulcrum's "stand" should be in the middle of the rocker arm slot when the valve is 50% open. Also, an imaginary line drawn across the top of the fulcrum and across the tip of the valve should be perpendicular to the valve stem’s axis when the valve is 50% open.

If nothing else:
• Verify that the fulcrum stands shall not bottom-out in the rocker arm slots at max lift.
• Verify that there are no push rod interference issues at or near max lift.
• Verify that there are no interference issues with the valve spring retainer when the valve is closed.

If adjustments are to be made to the fulcrum height, the same adjustments "should" be made for all 16 rocker arms in order to maintain the consistency of push rod length and tappet pre-load (assuming there even is consistency of push rod length and tappet pre-load). There is always the possibility that each head may be a little different.

If rocker arm geometry is going to be adjusted, it should be adjusted before establishing push rod length.

Camshaft Installation & Timing
Install a new replacement Q code (Cobra Jet) camshaft, or an aftermarket camshaft, and a new set of tappets. Aftermarket versions of the Q code camshaft are available from Melling # SYB-29, Sealed Power # CS-650 and Manley # MS700.

Torque the thrust plate bolts to 9-12 ft./lbs. The camshaft end play spec is 0.001-inch to 0.006-inch.

Install the Q code cam 4° advanced with a new multi-index timing set. DO NOT purchase a timing set without holes or slots in the cam sprocket. Refer back to comments in the short block section for an explanation. Torque the camshaft sprocket bolt to 40-45 ft./lbs. Don't forget the two piece fuel pump eccentric.

A factory CJ camshaft with 117° LSA, advanced 4°, should have the following timing factors:
• The seated opening of the intake valve (IVO) should occur at 18° BTDC measured at 0.006-inch tappet lift.
• The seated closing of the intake valve (IVC) should occur at 72° ABDC also measured at 0.006-inch tappet lift.
• The intake centerline (ICL) should occur at 117° ATDC.

An aftermarket CJ camshaft with 115° LSA, advanced 4°, should have the following timing factors:
• The seated opening of the intake valve (IVO) should occur at 20° BTDC measured at 0.006-inch tappet lift.
• The seated closing of the intake valve (IVC) should occur at 70° ABDC also measured at 0.006-inch tappet lift.
• The intake centerline (ICL) should occur at 115° ATDC.

A degree wheel, and a dial indicator are needed to verify camshaft timing.

Push Rod Length
Modern hydraulic tappets have 1/3 as much plunger compression as they once had. This lessens their ability to adjust for variations in dimensions. Between milling the heads, possibly using thinner (0.038) head gaskets, and using modern tappets, the original push rods will be too long. The new length will need to be determined, and replacements will need to be purchased.

To determine the length of new push rods, the heads shall need to be assembled using valve train checking springs. Use one head at a time, torqued in place with a "sacrificial head gasket"*. Also needed are a pair of "solid tappets"**, a pair of adjustable push rods (aka measuring push rods), and a pair of rocker arms. Check the length of the intake & exhaust push rods, at all 4 corners (cylinders 1, 4, 5, and 8), with both heads. The more consistent the measurements are, the better. A method to accurately measure the push rod lengths is needed, they are over 8 inches long.

*Note: A "sacrificial head gasket" is a third head gasket that shall be crushed to its compressed thickness during mock-up assembly. It should not be used during final assembly. Thus, when purchasing head gaskets, purchase at least three.

**Note: A "solid tappet" is a spare hydraulic tappet that has been disassembled and its internal parts replaced by a stack of washers. Thus, when purchasing tappets, purchase a couple of spares.

"Ideally" all 16 new push rods should be the same length, and all 16 tappets should be pre-loaded approximately the same. More realistically however,  variations in dimensions should be small enough to be within the tappets' ability to adjust for them.

The new push rods should be long enough to compress (pre-load) the tappets by 0.010-inch to 0.020-inch where the longest push rods were measured (at least 0.010-inch). That same push rod length should fall short of bottoming-out the tappet plungers where the shortest push rods were measured.

To gain rpm at the top and extend the rev limit the new push rods must be stiffer than the factory push rods. Purchase them made from seamless chromoly tubing, either 5/16-inch OD x 0.116-inch wall or 3/8-inch OD x 0.083-inch wall. Smith Bros. Push Rods of Redmond Oregon is one possible place to order custom length push rods.

Valve To Piston Clearance
If an aftermarket camshaft is being installed (higher lift, longer duration), the valve to piston clearance should be checked. It's better to be safe than sorry. A head with valve train checking springs, the sacrificial head gasket, a pair of "solid tappets", a pair of adjustable push rods, a pair of rocker arms, a degree wheel, and a dial indicator are needed to perform that check. Clearance checks should be performed from 15° BTDC to 15° ATDC during the overlap period. The recommended minimum intake valve to piston clearance is 0.080-inch. Floating exhaust valves are the valves that get hit by pistons. Some quote 0.100-inch minimum exhaust valve to piston clearance, others quote 0.120-inch.

Now that the valve to piston clearance has been checked, and any issues regarding valve to piston clearance have been resolved, there is no longer a  possibility that the reciprocating assembly may have to be pulled from the short block. Assembly of the short block can resume.

Frontal Assembly & Oil Pan
Temporarily install the new crankshaft damper and the timing pointer. Turn the crank so that cylinder #1 is set at 16° BTDC on the compression stroke (both valves closed) as per the timing marks on the crankshaft damper. Install the distributor so that a vane on the armature aligns with the metal strip in the center of the magnetic pick-up, and the rotor aligns with the #1  terminal on the distributor cap. Standing in front of the engine, looking down on the distributor, the no. 1 terminal is located at approximately 1 o'clock to 2 o'clock. Finally, the vacuum advance mechanism should point "more or less" straight ahead. It may be canted to the left just a little bit, like in the left hand diagram. This usually takes a few tries to get it right. Clamp it down in that position.

Dist Install

Re-use the oil pump or purchase a new standard volume pump; either way the oil pump should be taken apart, inspected, cleaned, pre-lubed, and reassembled with lock-tite on the bolt threads. The oil pump "pick-up" must be threaded and tightened into the pump prior to installing the pump. Torque the oil pump to block bolts to 25-35 ft./lbs.

DO NOT install a heavy-duty oil pump drive shaft. Use the stock oil pump drive shaft, or a stock replacement.

Install the fuel pump. The factory fuel pump was supplied by Carter. The part number for Carter's current replacement is # M6882. Torque the fuel pump bolts to 14-20 ft./lbs.

The timing cover is often badly corroded around the coolant passages, and under the coolant pump. New zinc plated timing covers are available from Summit Racing, # SES-5-65-04-201. Install a new seal in the timing cover, slip the oil slinger (deflector) on the crank snout, and install the cover. Torque the timing cover bolts to 14-20 ft./lbs. Don't forget the timing pointer.

The rubber sleeve of a 20-year-old crankshaft damper, hardened by age and heat cycling, no longer dampens as designed. The hardened rubber also allows un-bonded outer rings to “walk” on the hub; thus the timing marks are most likely incorrect. A 20-year-old damper is ready for replacement; an even older damper is long overdue for replacement.

Replace the crank damper with Powerbond's #PB1082SS damper. It is a heavy, fully bonded, steel replacement which is suitable for performance engines. The damper bolt torque spec for dampers with steel hubs is 130-150 ft./lbs. The damper bolt torque spec for the OEM damper (cast iron hub) was 70-90 ft./lbs.

Install the crank pulley too. The crank pulley bolts have two torque specs, torque "UBS bolts" to 25-35 ft./lbs, torque "place bolts" to 35-45 ft./lbs.

Pantera Cooling System

Install the coolant pump. The Milodon and Weiand coolant pumps have blocked "bypass" passages as they come out-of-the-box. The passages should be drilled open prior to installation, or a different pump should be chosen. The Pantera's cooling system performance is improved by a coolant pump designed to increase low rpm flow (Flow Kooler # 1648), and/or an over-drive coolant pump pulley (IPSCO # IPS260-OD). Torque the coolant pump bolts to 14-20 ft./lbs.

Racing Oil Pan

Install a wet sump racing oil pan. Choices include those manufactured by Armando, Kevko, Aviaid, and Pantera Performance Center (Dennis Quella). Torque the smaller oil pan bolts (1/4-inch) to 7-9 ft./lbs, torque the larger bolts (5/16-inch) to 11-13 ft./lbs.

Cylinder Head Assembly
It's time to finally remove the valve train checking springs and install the actual valve springs.

The factory M code or Q code valve springs are good parts, reuse them with the factory cam (1.82 installed height, 90 lbs. seated spring force, 390 lbs/in spring rate, 0.505-inch max lift). If the factory springs are reused the 7° retainers, rocker arm fulcrums, and rocker arms should be reused as well. The valve spring specs are:
Free length (approx.): 2.05-inch
Assembled height, pad to retainer: 1.81-inch to 1.84-inch
Maximum out of square: 0.078-inch
Minimum seated force: 79 lbs at 1.82-inch height
Minimum open force: 244 lbs at 1.32-inch height

Manley Performance sells a good replacement for the Q code valve spring, it's a Street Master valve spring # 22408-16 which is a double spring plus damper (1.82 installed height, 107 lbs. seated spring force, 392 lbs/in spring rate, 0.620-inch max lift). Use it with Manley # 23645-16 valve spring retainers (7°, standard height), and Manley # 42126-16 valve spring cups (0.062-inch thick). The spring cups are fitted to the head with cutter # 41835.

Heavy duty (hardened steel or heat treated) locks are needed to mate the 11/32 inch valve stems of the new valves (single groove or 4 groove) with the factory or Manley 7° retainers.

The M code/Q code rocker arm fulcrums were made of steel (sintered iron) whereas 2V rocker arm fulcrums were made of aluminum. The M code/Q code fulcrums, if needed, are available from Melling (# MRM-1776) or Sealed Power (# MR-1811). Assemble the rocker arms and fulcrums "finger tight" for now.

Final Assembly (finishing the long block)
The OEM head gaskets had 0.047-inch compressed thickness (10.2 cc). That is also the thickness of the popular Fel Pro # 8347 PT-1 composite head gaskets.

0.038-inch head gaskets have always been a part of raising the compression of a 351C. Ford’s five layer folded head gasket # D3ZZ-6051-A (sold over the counter) and McCord’s # 6850 head gasket had 0.038-inch compressed thickness (8.2 cc). Both gaskets were mentioned often in ‘70’s and ‘80’s literature. Edelbrock (# 7328) and Mr. Gasket (# 5808G) offer composite head gaskets in that thickness today.

Bolting the assembled 61cc heads on a standard bore short block with nominal 0.035-inch deck clearance and 3 cc piston dome volume, utilizing 0.038-inch-thick head gaskets, produces about 10.06:1 compression.

Cast over-bore pistons should compensate for an over-bore by reducing compression height OR increasing dome volume (slightly dished domes). If so then 0.038-inch head gaskets will be appropriate for an over-bore as well.

The following piston dimensions maintain the compression ratio for various bores:
• Std. 4.000 bore: 1.650-inch compression height and 3.0 cc dome volume
• 4.010 over-bore: 1.648-inch compression height or 3.4 cc dome volume
• 4.020 over-bore: 1.646-inch compression height or 3.8 cc dome volume
• 4.030 over-bore: 1.644-inch compression height or 4.2 cc dome volume
• 4.040 over-bore: 1.642-inch compression height or 4.6 cc dome volume

However, if over-bore pistons have the full 1.650 inch compression height AND only 3 cc dome volume then 0.047-inch head gaskets would be necessary.  This is unlikely if off-the-shelf cast pistons are employed.

Static compression and/or dynamic compression MUST be recalculated to account for alterations in deck height, alterations in compression height, alterations in piston dome volume, alterations in head gasket specs, alterations in combustion chamber volume, alterations in camshaft specs, or alterations in cam timing.

Fel Pro head gasket

The head gaskets have a definite front and back; take caution to orient them properly. Don't forget the head dowels either.

The factory torque spec for the head bolts stipulated torqueing them in 3 steps first to 55 ft./lbs, then to 75 ft./lbs, and finally to 95-105 ft./lbs. If a head bolt, head stud, or head gasket manufacturer stipulates different torque specs they should be followed.

Old torque specs stipulating 120 ft./lbs final torque were intended only for Ford's D3ZZ-6051-A head gasket, which was a unique 5 layer folded head gasket.

Assuming 10.06:1 static compression (nominal), and the seated closure of the intake valve (IVC) occurring between 66° and 72° ABDC, the dynamic compression (nominal) shall fall in the following range:

    66° IVC = 7.96:1 dynamic comp. ratio
    68° IVC = 7.83:1 dynamic comp. ratio
    70° IVC = 7.70:1 dynamic comp. ratio
    72° IVC = 7.57:1 dynamic comp. ratio

Once the heads are torqued in place, install the push rods and fasten down the rocker arms; torque the rocker arm fulcrum bolts (5/16-inch) to 18-25 ft./lbs.

4 bbl Carburetor
The emissions tuned factory Autolite/Motorcraft 4300 carburetor is not compatible with raised compression. It's too lean, using it would result in knocking, pinging, and detonation, so it must go.

Over the years Holley has released several 600 cfm 4bbl carburetors with vacuum secondaries; they were 4160 style carburetors (no secondary metering blocks), they had electric chokes, side hung floats and single feed fuel connections. They were ALL emissions tuned carburetors. The list 6619 was one common substitute for 351 Clevelands. Horsepower was reduced by 5% with that carburetor.

"That is not the type of carburetor you're looking for".

Not the Drones

This "de-smogged" high compression engine needs a "non-emissions" carburetor, tuned for about 12:1 A/F ratio, the kind that carries a warning "not for use on pollution-controlled vehicles".

The carburetor should be a square bore carburetor with venturis in all 4 barrels and dual high-capacity fuel bowls. That spec eliminates a bunch of carburetors. The carburetor should also have vacuum secondaries, and an electric choke. Annular booster venturis are cool, but only a "necessity" when the intake manifold is operated without exhaust heat.  Select the carburetor from those sold by Holley, Quick Fuel, or Summit Racing.

The carburetor should be rated 600 to 780 cfm the way Holley and Ford rates them. Since a 500 horsepower or a 7000-rpm engine is not being assembled, a smaller carburetor (600 to 650 cfm) shall support the engine adequately. For example, Ford's first high performance engine, the 352 High Performance of 1960, made 360 horsepower equipped with a Holley carburetor rated merely 540 cfm. Ford also installed 600 cfm Holley carburetors on 390 horsepower hydraulic-cammed (W code) 427s of 1968.

If a carburetor larger than 650 cfm is preferred, its actually too big for the application; but that's OK as long as it has vacuum secondaries. As an example of that, Ford installed 780 cfm Holley carburetors on 290 horsepower Boss 302s.

There's a lesson in there somewhere.

A 600 cfm Holley has 1.25-inch primary venturis, 1.31-inch secondary venturis, and 1.56-inch throttle bores. A 650 cfm Holley has the same size venturis but 1.69-inch throttle bores. Hopefully the smaller carburetor shall serve as somewhat of a rev limiter. The smaller carburetor may also improve the powerband width and therefore low speed performance. If I were in the market for a carburetor today Holley's 650 cfm vacuum secondary carburetor # 0-80783C is the one I would purchase. Check it out.

Mount the carburetor on the engine via a dual plane intake manifold, either the Ford iron 4 bbl manifold designed for 600 cfm Autolite 4300 square bore carburetors, or the Edelbrock Performer manifold (4V # 2665, 2V # 2750). The Edelbrock manifold becomes a Ford manifold if the Edelbrock logo is removed and it is painted Ford blue .

Machine Shop Task #5
The Ford manifold is a 4-hole type manifold and will require having the primary bores machined open to 1.69 inches diameter. And plug the heat passage in front of the primary holes.

The tin “turkey pan” intake manifold gasket is a necessity when the intake manifold is heated with exhaust heat. It shields oil splash from the “hot” exhaust crossover below the manifold, thus extending the life of the motor oil.

Torque the smaller intake manifold bolts (5/16-inch) to 23-25 ft./lbs, torque the larger bolts (3/8-inch) to 28-32 ft./lbs.

If the intake manifold is not heated by exhaust heat, then the carburetor should be equipped with annular booster venturis, so that fuel atomization within the manifold will occur similar to how it would occur if the manifold were heated. Carburetors can be custom ordered with annular boosters from both Holley and Quick Fuel, and they're standard equipment on Summit Racing's carburetors.

Not to be Forgotten
There are external components I haven't mentioned. The thermostat, the alternator, the starter, the valve covers, the spark plugs, and the ignition wires. Service or replace them as you see fit.

Cleveland thermostat

The 351C uses a special thermostat, a Robertshaw model 333. It is basically the more common Robertshaw model 330 thermostat with a flanged copper sleeve pressed on the bottom. There should be a brass orifice plate below the thermostat.

autolite racing

Older numbered spark plugs for an engine with 10:1 compression included Autolite AF32 (A = 14mm, F = conical seat), Autolite & Motorcraft ARF32 (A = 14mm, R = resistor, F = conical seat), and Motorcraft ASF32C (A = 14mm, S = suppressor, F = conical seat, C = copper core).

Newer spark plugs include Autolite 24, Autolite Racing AR24, or Motorcraft stock #15. Duraspark 1 plug gaps are 0.060-inch, Duraspark II plug gaps are 0.050-inch.

This freshened-up "351 Cobra Jet" has a potential output of 350 horsepower (320 horsepower with 2V heads), dependent upon how well the rings seal within their bores. More importantly, it would have 10:1 compression. That's where the pep, throttle response, acceleration, and low rpm torque come from.

The "de-smogged" engine's performance is a startling improvement over any "production" 351C, especially the Muskie Act compliant post-1971 versions. It's a good place to start, to figure out if going any further is even necessary.

Take It To The Limit
Add headers and tail pipes … perhaps the Hall GTS system … and the engine's output shall increase to 365 horsepower. Add a Blue Thunder manifold and the engine's output shall increase to 390 horsepower. Adding a camshaft with 0.520-inch net valve lift would nudge the output to about 425 horsepower.

A camshaft with 0.520-inch net lift would have required a long duration solid flat tappet camshaft in 1970, today that much lift can be achieved with a more street-able short duration, low overlap, hydraulic flat tappet camshaft (0.530-inch gross lift). That camshaft can be ordered, custom made, from Bullet Racing Cams specifying lobes # H274/3066 and #H282/3066,  specifying 114° LSA, and specifying the indexing (timing) as 5° advanced.

Advertised duration would be 274°/282°. Duration at 0.050-inch would be 222°/230°. Gross valve lift would be 0.530/0.530-inch. The LSA would be 114° and overlap would be 50°. The hydraulic intensity is 52°.

With 114° LSA, and indexed 5° advanced, this cam should have the following timing factors:
• The seated opening of the intake valve (IVO) should occur at 28° BTDC measured at 0.006-inch tappet lift.
• The seated closing of the intake valve (IVC) should occur at 66° ABDC also measured at 0.006-inch tappet lift.
• The mathematic intake centerline (ICL) should occur at 109° ATDC, but the lobes are asymmetric, thus max-lift shall occur several degrees earlier.

An off-the-shelf alternative is Crower's camshaft # 15966. Advertised duration is 278°/284°. Duration at 0.050-inch is 220°/226°. Gross valve lift is 0.529/0.540-inch. The LSA is 112° and overlap is 57°. The hydraulic intensity is 58°.

With 112° LSA, and indexed 4° advanced by Crower, this cam should have the following timing factors:
• The seated opening of the intake valve (IVO) should occur at 31° BTDC measured at 0.006-inch tappet lift.
• The seated closing of the intake valve (IVC) should occur at 67° ABDC also measured at 0.006-inch tappet lift.
• The intake centerline (ICL) and thus max-lift should occur at 108° ATDC.

Either cam would require valve springs compatible with hydraulic flat tappet cams having higher lift rates. Manley Performance has a Street Master valve spring # 22407-16 which fits the application. It is a double spring plus damper (1.82 installed height, 126 lbs. seated spring force, 418  lbs/in spring rate, 0.635-inch max lift). Use it with Manley # 23630-16 titanium valve spring retainers (10°, standard height), and Manley # 42126-16 valve spring cups (0.062-inch thick). The spring cups are fitted to the head with cutter # 41835.

Heavy duty (hardened steel or heat treated) locks are needed to mate 11/32 inch valve stems (single groove or 4 groove) with the Manley 10° retainers.

The # 22407 valve spring matched with a 0.530-inch lift hydraulic flat tappet cam fits in well with the flat tappet street cam limits formerly suggested by Crane Cams (seated spring force 115 to 130 lbs., spring force at max. lift ≅ 330 lbs., hydraulic intensity ≥52°).

The factory spring force with the original (hydraulic flat tappet) cobra jet camshaft was 277 pounds at 0.481 lift. The spring force at 0.530 lift with the Street Master valve spring shall be 347 pounds. The stiffer push rods suggested previously will be compatible with the stiffer springs and higher lift rate of this valve train. However, if the factory rocker arms are utilized, then install them with ARP 5/16-inch chromoly bolts # 641-1500 to eliminate the possibility of bolt stretch. If the bolts are too long, then shorten them or use 1/8-inch-thick ARP washers # 200-8587.

One of Ford's rocker arm suppliers manufactured rocker arms which had push rod interference issues at 0.550-inch valve lift. The rocker arms can be identified by the trapezoid shaped “bumps” along the edges of the rocker arm, located on either side of the fulcrum. This is not a problem for a valve train employing a factory camshaft. But with a 0.530-inch lift camshaft that would be cutting things close. In that situation those rocker arms should not be used, replacements shall be needed.



Images (18)
  • relief 2
  • Pantera Cooling System
  • Not the Drones
  • Bronze Guides & Positive Seals
  • Fel Pro head gasket
  • Racing Oil Pan
  • Cylinder Numbering
  • Cleveland thermostat
  • Pantera Cooling System
  • Crank Polishing
  • rocker geometry
  • rocker tip wear pattern
  • tappet bore bushings
  • Dist Install
  • autolite racing
  • 3 angle valve seats
  • Distributor Gear
  • Factory Rocker Arm Geometry
Last edited by George P
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?
the cast numbers on my block are
2K18 (date ?)
Yes, the date.

Engine block engineering number. Correct for the original build on this car.

stamped 01527 (I've lost the tag with engine number to verify)
This is the date the engine was assembled.Are you sure it isn't 01572? (January 5, 1972)

on the front inside corner of the heads is a "4"
The engineering ID for the 4v heads.

under the valve cover I only found
2J2 (date ?)
The casting date of the heads. It's rare to find them both exactly the same date.

circle with L5316
Escapes me at the moment

large C with F in center
The "mint mark" of the Ford Cleveland Foundry (casting facility)

would the head cast number be under the intake?
Yes. It is under one of the intake runners. You have to remove the head to see it. If the intake manifold is off you should be able to see it with a mirror.

Originally posted by JFB #05177:

... D2AECA ... corner of the heads is a "4" ... under the valve cover I only found 2J2 ...

All 1972 - 1974 blocks, and all blocks cast in Australia are D2AE-CA blocks. Both 2V and 4V. If it was originally a 4V block it should have 4 bolt main bearing caps. But ... don't ever let the lack of 4 bolt caps disappoint you about the block. The 2 bolt caps, when tightened properly, are stable to 8000 rpm! Don't ever let a mechanic talk you into aftermarket caps (steel, splayed bolts, etc.) they are not needed for the 351C. A set of main cap studs will help insure the caps are torqued correctly, at the very most that's all you need.

All 4V heads have a 4 in the corner, but only the 1970 heads have a "plain" 4, the 1971 - 1974 heads have a 4 and a "dot" in the corner. As Doug mentioned the head's casting number is actually underneath one of the intake runners, making it difficult to determine which head you have ... BUT ... you can check the casting's date code under the valve cover and get a good idea of which head it is. In your case 2J2 decodes as 1972 September 2nd. Because that head was cast so late in the year, its probably a head for a 1973 engine, meaning it has the largest chambers and the smaller 2V valves.

If the heads are original to the engine, and the engine has never been torn apart, then I expect it to be a 1973 Q code engine with dished pistons, 78cc combustion chambers and 8:1 compression. This would make sense since chassis number 5177 is a 1973 Pantera. The camshaft was ground 4° retarded. It was originally equipped with EGR. Installation of a set of Australian 302C heads (small port, 58 cc quench chambers) will raise the static compression to about 10:1 without doing anything else. Replace the factory induction with an Edelbrock Performer 2V intake manifold (#2750) and a 650 cfm Quick Fuel, Demon, or Holley "non-emissions" carb with vacuum secondaries; also advance the camshaft 4°. Those 3 changes alone shall increase the engine's pep and horsepower a significant amount.

Last edited by George P

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