Compact internal combustion engine

An engine which employs a cam follower mechanism to reduce wear and reduce the size of an assembled engine. The cam follower mechanism utilizes guide rails located to reduce side thrust on the valve stem. The engine employs a high speed quill shaft to synchronize independent cam shafts existing in each of a plurality of interconnected engines. The engine is assembled using a single size fastener to provide a uniform stress gradient within the engine. The engines are interconnected utilizing O-ring seals. The engine provides a piston crown utilizing a connecting rod directly connected to the bottom surface of the piston crown. The piston crown is stabilized along the longitudinal cylinder axis by a rail guide. Connecting rods are provided which require less than one hundred eighty degrees (180.degree.) circumference of a crankshaft pin for support so that a plurality of connecting rods can be associated with a single crankshaft pin. A tabbed bearing fits under the plurality of connecting rods to provide lubrication between the connecting rods and the crankshaft pin. Connecting rods are held to the crankshaft pin by a circular retaining ring. The engine provides a separate cylinder head and cylinder which are attached via a circular deformable retaining band to form a metal to metal seal. The engine provides an independent lubrication system in each engine. Coolant or lubricant is provided to each engine in parallel so that the temperature of the coolant entering each engine is the same. A large diameter modular crankshaft is provided.

BACKGROUND 
1. Field of the Invention 
The present invention relates to internal combustion reciprocating engines 
and in particular to a reduced size internal combustion reciprocating 
engine of which a plurality can be interconnected to form a larger engine. 
2. Description of the Related Art 
Internal combustion reciprocating engines have been known for over a 
century. The internal combustion reciprocating engine has been 
manufactured in numerous configurations over the years. These engines are 
utilized in automobiles, air planes and water craft. An important 
consideration in each of these applications is the size and weight of the 
engine. There is a trade off between the structural integrity or 
durability of an engine and the size and weight of the engine. Engine 
manufacturers design overly massive engine parts to increase the 
durability and useful life of an engine. Utilization of massive engine 
parts, however, increases the weight and size of the engine and can 
actually increase engine wear by increasing the dynamic weight of the 
moving parts in the engine. Thus there is a need for a reduced weight and 
size engine that is durable. 
Some engine manufacturers have apparently built engines by interconnecting 
a set of smaller engines or modular engines. Modular engines are known in 
the prior art as evidenced by the Voorhies patent, U.S. Pat. No. 
2,491,630, entitled "An Engine Constructed of Sections Bolted Together 
Along the Vertical Plane to Form an Entire Head Block and Crankcase 
Thereof," issued on Dec. 20, 1949. Voorhies patented an internal 
combustion engine constructed from a series of engine modules. The 
Voorhies engine however suffers the same inadequacies as other 
conventional engine designs. 
Some of the problems presented by typical engine designs are discussed 
below. 
Cam Followers 
Typical cam follower mechanisms act as an intermediary between a cam shaft 
lobe and a valve stem. Cam followers compensate for rotating cam lobes 
side thrust. Lobes assert a composite thrust containing both a horizontal 
(side thrust) and vertical (downward thrust) component. The cam followers 
absorbs some of the side thrust. Any portion of this horizontal thrust 
component which is asserted on the valve stem increases wear on the valve 
stem and valve stem guide in which the valve stem slides. The horizontal 
and vertical components are asserted upon the cam follower by the rotating 
cam lobe. The cam lobe rotates, depresses the cam follower mechanism, 
which in turn depresses the valve stem. Typically a portion of the side 
thrust component is not compensated for by the cam follower. This side 
thrust is asserted on the valve stem which increases wear on the valve 
stem and the valve stem guide. 
Typical engine designs typically provide minimal lubrication between the 
valve stem and the valve stem guide. Inadequate lubrication exacerbates 
the effect of wear caused by the side thrust asserted on a valve stem by 
the typical cam follower mechanism. Typically, engine designers utilize 
long valve stems to provide a relatively long longitudinal dimension, or 
high aspect ratio of length to width, in order to achieve stability of a 
valve stem along its axial length. 
Engine designers also consider the aspect ratio of the cup-type cam 
follower. The longitudinal dimension of a conventional cup-type cam 
follower assembly must be long enough to stabilize the cam follower along 
its axial length, therefore seeking to reduce the horizontal thrust 
exerted on the valve stem. As the cam lobe rotates and depresses the cup, 
the cup's resistance to the side thrust component is manifest in wear on 
the cup along a line 90.degree. from the axis of rotation of the cam lobe. 
In a typical cup-type cam follower, the top of the cup or cup face must 
have sufficient diameter to cover the valve spring. This cup 
configuration, thus requires a cup wide enough to cover a valve spring and 
long enough to be stable. The requirement for large cup increases the 
overall size of the assembled engine. 
Crankshafts 
Typical single piece and modular crankshafts have suffered harmonic 
breakage problems. These problems occur when the natural frequency of 
vibration of the modular crankshaft matches the frequency of impulses 
applied to the crankshaft, resulting in breakage, or can induce 
intolerable torsional deflections of the crankshaft. 
The typical high RPM engine produces power input pulses near the frequency 
range of the natural resonant frequency of the typical crankshaft. Thus, 
typical modular crankshafts tend to suffer from breakage as the input 
frequency matches the natural frequency of vibration. Typical modular and 
single piece crankshafts may also be distorted and strained from bending 
moments asserted on the crankshaft by the force of the pistons pushing the 
crankshaft pins. 
Cam Shafts 
Typical cam shaft deflection has caused typical engine designers to have 
problems synchronizing interconnected engine modules together to achieve 
appropriate timing. The cam shaft twists due to the twisting torque 
applied to it, adversely affecting the timing and the synchronization 
between engine modules. Typical engine designers utilize a large cam shaft 
to reduce twisting of the cam shaft in an attempt to overcome timing 
problems. Large typical cam shaft designs, however, increases the overall 
size and weight of the assembled engine. 
Engine Assembly 
Typical engine assembly utilizes a wide array of nuts, bolts and washers of 
varying shapes, sizes and lengths to assemble the parts to make a typical 
engine. The typical engine is assembled by different fasteners each having 
different torque requirements for each individual part of the engine. 
Different fasteners and different torque create a nonuniform stress 
gradient on the typical assembled engine. Nonuniform stress distorts the 
shape of the engine. Diversity of fasteners creates inventory overhead 
work for the engine manufacturer. The manufacturer must keep up with a 
wide variety of different size nuts and bolts. Thus, a wide variety of 
tools are required. Typical engines are assembled utilizing a different 
tool and assembly procedure for each part of the engine. Typical engines 
also utilize gaskets between metal parts which creates an assembled 
tolerance variation. Gaskets variably compress to a nonuniform thicknesses 
according to the pressure applied to the gasket. The pressure varies at 
each fastener and at each fastener location. Thus the tolerance of the 
assembled engine can vary as the thickness of the sealing gaskets vary. 
When assembling modular engines designers have found that typical engines 
require a different size oil pump and cooling pump for each different 
modular engine configuration, depending upon the number of modules 
connected to construct the engine. Oil pump size varies with engine size. 
Thus, the manufacturer must supply a different size coolant and 
lubrication pump for each configuration of one, two, three, four, or five 
typical engine modules connected together to construct an engine. 
Typically lubrication and coolant fluid flow serially through 
interconnected engine modules so that the lubricant and coolant fluid 
enter the first engine module where the fluid is pre-heated by the first 
engine module before the fluid enters the second engine module, third 
module, fourth module, and so on. Thus, the fluid entering the last engine 
module is substantially warmer than the fluid that entered the first 
engine module. Thus each typical interconnected engine modules run at a 
different temperature. 
Pistons 
Typical piston assemblies utilize a trunk style piston. The trunk piston 
has a flat circular top and a long cylindrical body or trunk. The trunk of 
the conventional piston fits closely within the cylinder. The cylinder 
wall guides the trunk of the piston and provides for stability of the 
piston along the longitudinal axis of the cylinder. The trunk of the 
conventional piston must be long enough, relative to the diameter of the 
piston, to provide adequate stability. The ratio of the piston length over 
the piston diameter determines how stable the motion of the piston is. The 
trunk of the piston rubs along the cylinder wall. The cylinder wall guides 
the piston. The additional weight of the elongated piston trunk increases 
the dynamic weight of the piston, thereby increasing the accelerative 
forces exerted on the piston, connecting rod and crankshaft pin. 
Typical pistons such as the trunk type piston, increase the overall size of 
the engine because the length of the cylinder must accommodate the 
additional length of the conventional piston trunk plus the displacement 
of the connecting rods. The typical trunk type piston also suffers from 
thermal expansion problems. Metal expands when heated. The trunk type 
piston swells to a large diameter when heated. Thus, the cylinder must be 
large enough to allow passage of the enlarged heated piston. The cylinder 
diameter must be large enough to maintain a substantial clearance between 
the cylinder wall and the piston trunk over the full range of engine 
operating temperatures. The clearance between the outside diameter of the 
conventional trunk type piston and the internal diameter of the cylinder 
wall must be maintained at all operating temperatures or the piston will 
"seize up" in the cylinder. Thus, typically, a substantial gap exists 
between the piston trunk and the cylinder wall to allow for variations in 
the diameter of the piston over the full operating temperature range of 
the engine. This excess gap left to allow for swelling of the piston 
creates a problem. At lower temperatures, there is a large gap between the 
piston trunk and the cylinder wall. At higher temperatures, the 
gap-between the piston and the piston wall is very narrow. The gap between 
the cylinder wall and the piston trunk, varies widely over the operating 
range of the engine. Thus there is a variation in the stability of the 
piston along the longitudinal axis of the cylinder. 
These thermal expansion considerations require engine manufacturers to 
design within close tolerances yet leave large gaps to account for wide 
variations in piston size over the operating temperature range. Piston 
stability along the longitudinal axis of the cylinder varies widely over 
the operating temperature range. Moreover, high tolerance requirements 
slow down the manufacturing process, to insure that the high tolerance is 
maintained. Slower manufacturing, requires additional man hours and time 
to produce the engine. 
Connecting Rods 
Typically connecting rods encircle and rotate around a crankshaft pin. The 
connecting rod end which attaches to the crankshaft pin must be a certain 
minimum width so that adequate lubrication can be established between the 
connecting rod end and the rotating crankshaft pin. Lubrication is in 
adequate below this minimum width causing increased wear and mechanical 
failure. 
Typically engines utilize connecting rods which are open at one end and 
bolted to a semi circular connecting rod bracket to form a circle around a 
crankshaft pin. The two piece, nut and bolt connecting rod configuration 
requires considerable additional mass for the nuts and bolts, thereby 
increasing the dynamic weight and forces experienced by the crankshaft and 
connecting rod attached thereto. 
The typical connecting rod requires considerable space. Although some 
engines attach more than one connecting rod to each crankshaft pin, 
typically the rods are side by side on a single crank pin. In this 
configuration, each connecting rod applies a sheer force across the entire 
crank pin length, a distance equal to twice the width of the connecting 
rod at the crank pin. The sheer force and attendant bending moment can 
cause bending and even breaking of the crankshaft pin. 
Cylinder Head Seal 
Some typical engines utilize a single piece head and cylinder assembly 
comprising a one-piece cylinder and cylinder head. This one-piece 
configuration presents a problem in machining the cylinder head. Machine 
bits must extend through the length of the cylinder to reach the machine 
surfaces of the attached cylinder head. Thus longer cutting bits must be 
used to reach the head. Longer bits are less rigid and thus reduce the 
accuracy of the head machining process. 
Other engines utilize a separate cylinder and cylinder head. Engine 
assemblers seal the cylinder head to the cylinder formed in an engine 
block with large bolts and gaskets. Gaskets are subject to variable 
thickness, depending upon the amount of pressure applied at each bolt 
location which the gasket seals. Irregular tolerances in an assembled 
engine decreases the structural integrity of the assembled engine. For 
example, typically, head bolt assembly methods rely on high pressures at 
isolated fastener points which deforms the engine block and degredates the 
structural integrity of the engine. Typical head sealing methods require a 
complex bolt tightening pattern to exact torque requirements. Such a 
methodology is prone to irregular assembly. 
SUMMARY OF INVENTION 
In accordance with the present invention, an engine is provided comprising 
one, or a plurality piston cylinders. A larger engine can be constructed 
from a plurality of the engines by interconnecting engines. Interconnected 
engine modules are sealed utilizing an O-ring. The engine provided by the 
present invention may be assembled and interconnected with a plurality of 
engines utilizing a single size uniform fastener. 
In accordance with the present invention, a modular crankshaft is provided 
having a crank pin comprising male and female portions. The male and 
female portions interconnect to form a crank pin. The connections also 
link crankshaft sections together. The male and female sections are 
splined together. 
In accordance with the present invention, a piston is provided comprising a 
piston having a crown. A rail guide assembly is attached to the bottom of 
the piston crown. The piston rail guide assembly rides on guides formed on 
the engine block in which the piston resides. The piston rail guide 
assembly stabilizes the piston crown so that the piston crown face remains 
perpendicular to the longitudinal axis of the cylinder in which it 
reciprocates. The piston is substantially smaller than the cylinder in 
which it resides which reduces wear on the cylinder wall. The piston crown 
center is guided along the center of the cylinder by thrust pads. Thrust 
pads attached to the bottom of the piston crown slide along the cylinder 
wall to guide the center of the piston crown within the center of the 
cylinder. 
In accordance with the present invention, a connecting rod is provided 
which at one end fits around a crankshaft pin and at the other end 
attaches to the bottom of the piston crown. The connecting rod does not 
fully encircle the crankshaft pin so that a plurality of connecting rods 
are held in place by a retaining ring which encircles a single crankshaft 
pin within the width of a single connecting rod. Connecting a plurality of 
connecting rods within the width of a single connecting rod on a single 
crankshaft pin, shortens the overall length of the crankshaft. A shorter 
crankshaft suffers less distortion during operation. 
The other end or small end of the connecting rod is rotatably attached to 
the bottom of the piston crown. The piston crown rail guide fits over and 
retains the small end of the connecting rod and a connecting pin. The 
connecting rods rotate about the connecting pin which abuts the bottom 
surface of the piston crown. The rail guide fits over and retains the 
connecting rod and pin under the piston crown. The connecting rod assembly 
shortens the overall dimensions of an engine and reduces wear on the 
connecting pin. 
In accordance with the present invention, a cam shaft is provided. The 
present invention provides a quill shaft which synchronizes the timing of 
separate and independent cam shafts which are provided in each of the 
separate but interconnected engines. 
In accordance with the present invention, a lubrication and cooling system 
is provided within each engine. Thus, a series of interconnected engines 
are inherently equipped with an appropriate lubrication and cooling 
system. In accordance with the present invention a valve head is provided 
which fits into the halves of an engine module. These and other provisions 
of the present invention are illustrated in the following description. 
The engine of the present invention provides a plurality which maybe 
duplicated to provide identical compact engines which may be 
interconnected to form a larger engine. Each engine contains either one, 
two, three or more cylinders. An eight cylinder engine can be constructed 
by interconnecting two four-cylinder engines or by interconnecting four 
two-cylinder engines. 
The engines are easily interconnected in metal to metal contact utilizing 
uniform fasteners and O-rings to form seals between interconnected 
modules. The uniform fastener reduces assembly time and helps to 
standardize assembly tools and methods. The modular engine uses a 
plurality of identical fasteners to assemble the entire engine. 
The engine of the present invention provides a cam follower apparatus that 
is configured to reduce the overall size of an engine while greatly 
increasing the allowable margin of error during the manufacturing process. 
Guide rails are provided on the cam follower body which attenuate the 
horizontal side thrust component of the cam lobe thrust, so that the valve 
stem is actuated essentially by only the vertical thrust which acts 
parallel to the valve stem's longitudinal axis of motion, reducing wear. 
The engine of the present invention provides a cam shaft in each engine 
module. The cam shaft in each engine module is synchronized with the cam 
shafts in other interconnected engine modules by use of an external high 
RPM quill shaft. The cam shafts are geared to the high speed quill shaft 
which reduces timing errors induced by twisting of the cam shaft. 
The engine of the present invention is assembled utilizing a plurality of 
uniform fasteners. Using a single fastener reduces the manufacture's 
requirement for inventorying of different size and length nuts and bolts. 
Uniform fasteners also simplify engine assembly methods. The uniform 
fastener enables the present invention to utilize a large number of 
uniform fasteners which evenly distribute the forces applied to the engine 
across the engine structure. 
The engine of the present invention provides a piston crown which utilizes 
thrust pads to center the piston crown within the center of a cylinder. 
Guide rails which run within guide slots are attached to the piston crown. 
These guide rails keep the piston crown face stable along the longitudinal 
axis of the cylinder. The stabilizing influence of the piston guide rails 
eliminates the need for the long piston trunk typically used in engines. 
The piston enables an engine manufacturer to assemble an engine which is 
smaller than a typical engine with the same stroke. This present invention 
provides a structure which reduces or eliminates the bending moment of the 
shear force acting on the connecting rod pin. Thus, the size and weight 
requirements for the connecting rod pin is reduced. The reduced size and 
weight of the pin connecting rod assembly reduces the dynamic weight and 
wear on the pin during operation of the piston assembly. 
The placement of the connecting rod abutting the lower surface of the 
piston crown enables the connecting rod to pivot close to the piston crown 
upper face. This configuration shortens the distance between the 
connecting rod end and the piston crown upper surface, which provides an 
engine smaller than a typical engine with the same stroke. Thrust pads are 
utilized to maintain the piston crown within the center of the cylinder. 
In the engine of the present invention, a plurality of connecting rod ends 
are connected to a crankshaft pin. The connecting rod end has 
substantially the same diameter and radius curvature as the crankshaft 
pin. A circular bearing between the crankshaft pin and the connecting rod 
end facilitates lubrication. A set of retaining rings is provided to 
maintain contact between the crankshaft pin and the connecting rod end 
assembly. 
Attaching more than one connecting rod end to a single crankshaft pin 
reduces the overall length of the crankshaft, which reduces the bending 
moment of the shear forces applied to the crankshaft by the pistons 
through the connecting rods. Reducing the bending moments induced in the 
crankshaft pins, by reducing their length overall, increases the 
structural integrity of the crankshaft during operation. The crankshaft is 
shorter than a typical crankshaft. 
The present invention provides a large diameter crank pins and crankshaft 
to reduce twisting and torsional deflections induced in the crankshaft. A 
tab on the connecting rod bearing restricts the rotational motion of the 
bearing relative to the connecting rod ends so that oil supply apertures 
in the tabbed bearing are not exposed to the gaps between the connecting 
rod ends. 
Cylinder Head Seal 
The cylinder head of the present invention is configured to facilitate 
machining of the intake ports, exhaust ports and valve guides in the 
cylinder head. The top of the cylinder is cut at an angle so that the line 
at the top edge of the angled cylinder edge creates a high loading when 
pressure is applied. This enables the angled cylinder edge to form a metal 
to metal seal against the cylinder head. A circumferential land around the 
cylinder, circumferential land around the cylinder head and a retaining 
band are utilized to attach and seal the cylinder to the cylinder head. 
the retaining ring and lands fit into a receiving grove cut in each engine 
block half. 
Lubrication System 
The engine of the present invention provides an independent lubrication 
system for each engine. Each engine contains its own independent 
lubrication and cooling system comprising a coolant pump, a scavenger 
pump, and a pressure pump. The oil supply is manifolded in parallel to 
each engine so that each engine is supplied with oil at the same 
temperature. Each engine module runs at the same temperature. A plurality 
of modules connected together to form an extended modular engine will have 
an appropriate lubrication system. A main supply line from the oil 
radiator outlet is manifolded in parallel through a constant temperature 
line into each of the engines so that the temperature of the oil at each 
engine is the same. 
Crankshaft 
The crankshaft is comprised of a plurality of modules which interconnect in 
a male-female fashion to form a crankshaft. The male-female crankshaft 
connections are splined together for rotational stability. The crankshaft 
is made of a stiff material with a large diameter so that the natural 
frequency of vibration of the crankshaft is much higher than the frequency 
of the rotational impulses applied to the crankshaft by the low RPM 
engine. Thus, the frequency of piston impulses does not enter the range of 
the crankshaft's natural frequency of vibration. This substantially 
reduces the probability of harmonic breakage problems due to piston 
impulses matching the natural frequency of vibration in a crankshaft. 
Valves 
The cylinder head of the present invention uses three intake and three 
exhaust valves for each cylinder. 
Other advantages and features of the invention will be apparent after 
studying the following description of a preferred embodiment.

DESCRIPTION OF AN EXAMPLE OF A PREFERRED EMBODIMENT OF THE INVENTION 
Turning now to FIG. 1, in the present example of a preferred embodiment of 
the invention, the engine of the present invention provides a plurality of 
identical engines which may be interconnected to form a larger engine. As 
shown in FIG. 1, each engine contains either one 10, two 12, three 14 
cylinders or more. Each cylinder houses a piston, e.g. 16, 18, 20. As 
shown in FIG. 4, these individual engines may be interconnected by 
abutting the planer surfaces 25 located mid-way 24 between the axial 
separation of the cylinders in adjacent modules. Each engine provides one 
piston 16, two pistons 18, or three pistons 20. The engines run 
independently or may be interconnected to work in cooperation. 
An eight cylinder engine can be constructed by interconnecting four 
two-cylinder modules 12 or eight one-cylinder modules 10. As shown in FIG. 
4, the engines are interconnected utilizing metal to metal contact at the 
axial plane 24 mid-way between adjacent cylinders 22. A metal to metal 
contact is formed between the adjacent planer surfaces 26 utilizing 
uniform fasteners discussed below. An O-ring groove is fashioned in the 
planer surface 26 of each engine. The O-ring is placed in the O-ring 
groove to form an O-ring seal between adjacent interconnected engines. 
As shown in FIG. 5, each engine is split into two halves 26 on a plane 
perpendicular to the longitudinal axis of the crankshaft 30. A groove 195 
is formed on the interior wall 147. A cylinder head 138 and cylinder 141 
are fastened together with a retaining ring 142 that fits into the groove 
195 and is secured when the two halves 26 are brought together (see FIGS. 
22, 23A, 23B, and 24. In the present example of a preferred embodiment, 
the engine utilizes a piston crown and piston guide rail assembly, rather 
than a trunk type piston. The piston crown assembly enables a designer to 
reduce the size of the engine and prolongs engine life by reducing induced 
wear. The piston crown assembly is stabilized by guide rails and thrust 
pads instead of the piston trunk. 
Referring to FIG. 6, the cylinder spacing within an engine is configured so 
that an engine can be intermeshed with an adjacent engine. The split plane 
of one engine becomes the separation plane between the intermeshing 
engines. Single cylinder engines can be interlaced into 2 cylinders, 4 
cylinders, 6 cylinders, etc., configurations. Three cylinder configure 
engines can be similarly interlaced as 3 cylinders, 6 cylinders, 9 
cylinders, 12 cylinders, etc., engine configurations. In the case of the 
meshed configuration engine, an extra crank throw is introduced between 
bearings. All other interfaces remain identical, differing only in axial 
dimension. 
In the present example of a preferred embodiment of the present invention, 
each engine provides a lubrication and coolant system and a cam follower 
apparatus. When a plurality of engines are interconnected, it becomes 
desirable to synchronize the firing of the pistons in the individual 
engines. Synchronizing enables proper timing of the overall composite 
engine composed of a plurality of engines running in synchronization. 
Therefore, the individual cam shafts in each engine are synchronized. In 
the present example of a preferred embodiment of the present invention, 
synchronization between the plurality of engines interconnected is 
facilitated by an external high RPM quill shaft, discussed below. 
In the present example of a preferred embodiment of the present invention, 
the engine utilizes a guide slot to stabilize the piston and guide rails 
to stabilize the cam follower mechanisms along their respective axis of 
translation. The guide slots and rail guides of the present invention are 
compact and require less space to perform their respective function than 
typical equivalents. Compact design for the guide slots and rail guides 
reduce the overall size of the engine and prolong its useful life. 
The entire engine can be assembled and interconnected with other modules 
utilizing a single uniform fastener and tool, discussed below. In the 
present example a preferred embodiment of the present invention, the 
engine utilizes a large diameter modular crankshaft, discussed below. In a 
preferred embodiment of the present invention, the modular engine is 
assembled utilizing a uniform fastener of constant size and length. The 
fastener positions 200 present in an engine are illustrated in FIGS. 27. 
The engines are connected in metal to metal contact providing a uniform 
cumulative assembled tolerance for the final assembled engine. Uniform 
cumulative assembled tolerance enables an engine manufacturer to 
interconnect a plurality of engines without experiencing cumulative 
tolerance errors between the engines. Cumulative tolerance errors may be 
experienced when a series of engines are interconnected with gaskets whose 
thickness may vary according to the force applied. The cumulative error 
experienced when gaskets are used, may become significant when 
interconnecting a stack of engines such eight two-cylinder engines, which 
could be interconnected to form a sixteen cylinder (V-16) engine. 
Cumulative tolerance errors may cause the engines to align improperly with 
the crankshaft, due to a variation in the longitudinal axis of the 
crankshaft. The metal to metal contacts of the present invention enable 
the eight engine stack for example, a V-16 to be uniform along the 
longitudinal axis of the crankshaft, without variations caused by the 
cumulative tolerance errors which may be caused by assembling with 
gaskets. 
In the present example of a preferred embodiment of the present invention, 
the engine utilizes the entire facial cross section of an engine to form a 
metal to metal contact, and O-ring form a seal between the entire facial 
cross sections of adjacent engines. Unlike the Voorhies modular engine 
discussed earlier, the engine of the present example of a preferred 
embodiment provides for metal to metal contact between entire cross 
sections of adjacent engines, enabling the present invention to achieve a 
more compact design along the longitudinal axis, that is, build a shorter 
engine. 
Cam Follower 
Turning now to FIG. 7, in the present example a preferred embodiment of the 
present invention a cam follower 32 is utilized to reduce the overall size 
of the engine and increase its useful life. As shown in FIG. 7B, cam 
follower mechanism 32 acts as a mechanical intermediary between the 
rotating cam shaft lobe 40 and the valve stem 54. As shown in FIG. 7C, cam 
shaft lobe 40 rotates about cam shaft axis 46. As shown in FIG. 7D, the 
guide rails 34 of cam follower 32 slide up and down in guide slots 37. 
Guide rails 34 are formed on the sides of the cam follower 32. Guide slots 
37 are formed in the cylinder head 35 and a cylinder filler block 33, 
which is installed or formed in the cylinder head. 
As shown in FIG. 7E, the thrust from the rotating cam lobe 40 may be 
resolved into a horizonal component 39 (side thrust) and a vertical 
component 41 (down thrust). The cam follower guide rails 34 absorb the 
side thrust component 39. Thus, only the down thrust cam lobe thrust 
component 41 is transmitted through the cam follower mechanism 32 to the 
valve stem 54. Reduction of side thrust reduces wear on a valve stem, for 
example, valve stem 54 and any associated valve guide. 
Guide rails 34 are utilized in the present example of a preferred 
embodiment to absorb the side thrust component 39 and to provide 
stabilization of the cam follower 32 along the axis of translation. The 
cam follower slides up and down on an axis parallel to the longitudinal 
axis of the cam follower guide rails 34. 
The cam follower reduces wear on the valve stem by attenuating the side 
thrust component 39 of the cam lobe thrust. Thus, only vertical thrust, 
parallel to the longitudinal axis of the valve stem, is asserted on the 
valve stem reducing wear thereon. Side thrust increases wear on the valve 
stem and thus reduces engine life. The cam follower mechanism 32 of the 
present invention operates in an oil lubricated environment within the 
cylinder head. 
Unlike typical cup-type cam follower mechanisms, as shown in FIG. 8, the 
present invention relies on the aspect ratio of the cam follower guide 
rail 34 rather than the aspect ratio of the diameter of the conventional 
cup-type cam follower 48. The cup-type cam follower 48 relies on its 
cup-shape for stability. The cup acts as a mechanical intermediary between 
the cam lobe 40 and a valve spring 50. Cam lobe 40 rotates about cam shaft 
axis 46. Cam lobe 40 depresses cup-type cam follower 48, which in turn 
depresses valve stem 54. Valve stem 54 is depressed along the longitudinal 
axis of the valve stem 54, and guided by valve guide 52. The conventional 
cup-type cam follower 48 relies on the aspect ratio defined by the 
diameter of the cup over the length of the cup, to achieve stability of 
the cam follower along the longitudinal axis of translation of the valve 
stem 54. The diameter 43 of the cup-type cam follower 48 must be large 
enough so that it will fit over the valve spring 50, or some other valve 
return mechanism. Therefore, the minimum diameter 43 of a cup-type cam 
follower must be slightly larger than the diameter of the valve spring 50. 
Thus, the diameter of the valve spring dictates the minimum length of side 
49 of the cup required to stabilize the cup. The large minimum diameter 
cup dictates a long minimum cup length, which increases the size of the 
engine. Typical engine designs utilize long valve stems to increase the 
aspect ratio of the valve stem and reduce engine wear. Long stems increase 
the overall size of the engine. The engine of the present invention 
provides compact short stem valve and valve stem. 
Referring back now to FIG. 7A, in the present example of a preferred 
embodiment of the present invention, the engine provides cam follower 32. 
Cam follower 32 relies on the aspect ratio of the guide rail 34 to absorb 
the side thrust and to achieve stability along the longitudinal axis of 
the valve stem. Cam follower 32 of the present invention does not have to 
fit over the valve spring as does the typical cup-type cam follower. This 
enables the cam follower of the present invention to provide a compact cam 
follower which reduces the required size of the cam follower and thus 
reduces the overall size of the engine in which it is embodied. 
Cam follower 32 of the present invention relies on the aspect ratio of cam 
follower rail 34 (the ratio of the length divided by the width of cam 
follower rail 34) for stability. Cam follower guide rail width is 
significantly less than that required in a cup-type cam follower, which 
must fit over the valve spring. The cam follower guide rail of the present 
invention does not have to fit over the valve spring and therefore is much 
smaller. Because the width of the cam follower guide rail 34 is 
significantly less than the required diameter of the cup-type cam 
follower, the cam follower of the present invention enables construction 
of a structure which provides high aspect ratio for the cam follower guide 
rail, yet utilizes significantly less space for any given aspect ratio. 
To maximize the guide rail aspect ratio, and the stability of the guide 
rail 34 the end 42 of the guild rail 34, which engages the cam shaft 43, 
as shown in FIG. 7B, is cut out to match the diameter of the cam shaft 43, 
which it engages. This maximizes the length of the face of guide rail 34 
adjacent guide rail slot 37. The longer rail length absorbs more side 
thrust and provides more stability to the cam follower along the cam 
follower's axis of translation. Thus, the cam follower is small but 
effectively attenuates the side thrust of the cam lobe. FIG. 9 is a 
detailed illustration of the cam follower guide rail 34, interface 42, and 
the cam shaft 
In the present example, the modular engine of the present invention runs at 
approximately 2700 RPM. The cam shaft RPM is approximately 1350. The lower 
RPM and compact design cam shaft reduces the accelerative forces asserted 
on the cam shaft, the cam follower and the valve assembly. Thus, the cam 
shaft can be easily constructed by pressing cam lobes onto the cam shaft 
to obtain an elastic fit, rather than using typical slower manufacturing 
methods which utilize a plastic fit. The reduced accelerative forces 
enable the engine to provide a compact and low pressure valve/valve spring 
apparatus. Thus, the engine provides a smaller diameter valve face, and a 
shorter length valve stem than typical valves. This compact design valve 
substantially reduces the dynamic mass of the valve of the present 
invention over that of typical prior art valve assemblies. 
Typical valves are long in order to enhance the stability along its axis of 
translation. The cam follower of the present invention efficiently absorbs 
the side thrust component of the cam lobe thrust so that less longitudinal 
stability compensation is required by the valve stem. Thus, the valve 
stems do not have to be as long because they do not have to compensate for 
instability, as are required by the typical valve stems. Thus the present 
invention valve reduces the required overall size of the engine. 
In the present example of a preferred embodiment, the engine utilizes six 
valves per engine. Utilizing six valves and a low RPM creates a very light 
valve requirement and with low inertia. Cam lobes can thus be stamped from 
sheet metal or made as powered metal pressings and pressed onto the cam 
shafts as shown in FIGS. 10A and 10B. 
Quill Shaft 
Turning now to FIG. 11, in the present example of a preferred embodiment, 
each engine 10 provides two cam shafts 56 and 58. Each cam shaft provides 
three lobes 60. The rotation of cam shafts 56 and 58 is synchronized by 
gear 62. FIG. 12 illustrates a series of interconnected engines 10. The 
timing of the cam shafts 56, 58 for each module is synchronized to the 
timing of the cam shafts in other interconnected modules. 
As shown in FIG. 12, in the present example of a preferred embodiment, the 
present invention utilizes a quill shaft 64 to synchronize the plurality 
of cam shafts 56 and 58 provided by each interconnected engine. The quill 
shaft is driven by step-up drive 66, which is attached to and driven by 
the crankshaft 68. The step-up drive 66 enables the quill shaft 64 to run 
at substantially higher RPM than the crankshaft. 
In the present example of the preferred embodiment of the present 
invention, the quill shaft RPM is twelve times that of the crankshaft. 
Quill shaft 64 comprises a plurality of interconnected sections 65. Each 
individual quill shaft section 65 is coupled to an individual engine cam 
shaft. The high RPM quill shaft reduces the torque for a given applied 
force exerted on the quill shaft. The torque exerted on the quill shaft is 
reduced by a factor of twelve or the ratio of the quill shaft RPM divided 
by the crankshaft RPM. The reduced torque induces less torsional 
deflection or twisting for a given horse power input, than it would at a 
lower RPM and the same applied horse power. Thus timing errors being 
induced by torsional deflections are significantly reduced or eliminated 
by the reduced torque, high RPM quill shaft. 
The quill shaft 64 of the present invention enables selection of a variable 
quill shaft size to accommodate a specified tolerable torsional 
deflection, or timing error, for an engine comprised of a given number of 
interconnected engines. Each individual engine is identical, thus each 
engine provides the same valves, crankshafts, cam shafts and cam lobes. 
The external quill shaft enables the engine designer to use identical 
engines to build up larger engines, and maintain independent control over 
timing errors between the engines by introducing a quill shaft to 
synchronize the timing between the engines. 
Piston 
The piston of the present invention enables the manufacture to assemble a 
engine which is smaller than a typical engine having the same stroke. 
Because the connecting rod is attached near the piston face at the lower 
surface of the piston crown, the engine cylinder length need accommodate 
only the stroke or axial displacement of the piston, without providing the 
additional length necessary to accommodate the trunk of a typical piston. 
The preferred embodiment of the piston assembly provides a smaller 
connecting pin than a typical piston. The engine enables a smaller pin to 
be utilized by reducing stress forces on it. The smaller pin reduces the 
dynamic weight of the piston assembly and, the associated accelerative 
forces asserted on it, thus reduces the connecting rod, and the crankshaft 
to which it attaches. 
Placement of the connecting rod abutting the lower surface of the piston 
crown enables the connecting rod to be attached close to the upper face of 
the piston crown, thereby shortening the distance between the connecting 
rod end and the piston crown. In the typical engine design, the distance 
between the piston face and the connecting rod end is increased by the 
length of the piston trunk. Thus, the piston of the present invention does 
not require as much space to accommodate the same engine stroke because 
the present invention does not have to accommodate the additional length 
of the piston cylinder trunk. 
The piston crown of the present invention seals the combustion chamber 
utilizing a piston ring. The piston crown does not rub against the 
cylinder walls. The piston crown utilizes thrust pads to slide along the 
cylinder wall guiding the center of the piston. The piston crown can be 
made of a material which expands and contracts readily under the varying 
temperatures experienced during engine operation. When the engine first 
starts, it is cold and the gap between the cylinder wall and the piston is 
relatively large. The piston crown contracts and expands. The crown is 
made of thermal conductive material which disburses heat without 
concerning the engine designer with the clearance between the piston crown 
edge and the cylinder wall. 
Turning now to FIGS. 13A-13C, the stability of the piston crown face as 
perpendicular to the longitudinal axis of the cylinder is provided by the 
cross head rail guide assembly, rather than the typical piston trunk. The 
present invention provides guide rails 84 and keys 82 to stabilize the 
piston crown face perpendicular to the longitudinal axis of the cylinder 
as shown in FIGS. 13A, 13B, 13C and 14. 
The stability of the piston face is dependant upon the aspect ratio of the 
stabilizing member. Typically, the piston trunk must be long enough 
relative to the diameter of the piston face to obtain a suitable aspect 
ratio and associated stability of the piston face with respect to the 
longitudinal axis of the cylinder. The present invention utilizes a cross 
head guide rail assembly to provide stability to the piston crown face in 
a plane perpendicular to the longitudinal axis of the cylinder. Thus, it 
is the dimensions of the small rail guide rather than the larger piston 
diameter which dictate the aspect ratio and stability of the piston crown 
in the present invention. The present invention provides greater stability 
utilizing a smaller space, because the stability of the piston crown in 
the present invention depends on the dimensions of the guide rails 84 
rather than the dimensions of the piston. 
The stability of a typical piston varies over the operating range of the 
engine, because the clearance between the stabilizing member, the piston 
trunk, and the cylinder wall, varies as the piston expands and contracts 
under temperature variations. The typical engine designer must allow 
sufficient space between the piston trunk and the cylinder wall to 
accommodate the expanded piston when hot and swollen. At cooler 
temperatures, the cooler piston has a smaller diameter. There is a larger 
gap between the cylinder wall and the piston trunk. Thus, there is less 
stability of the piston at lower temperatures when the piston cools than 
when it is hot. The stability of the typical piston varies over the 
operating temperature, as the gap between the piston trunk and the 
cylinder wall varies, when the piston expands and contracts. 
The rail guide assembly of the present invention maintains a much more 
consistent stability over varying temperatures. The rail guide is less 
sensitive to temperature variations. The rail guide relies on the smaller 
dimensions of the guide rail and associated guide key to maintain 
tolerances over a wide thermal and stable temperature of the piston. The 
width of the guide rail, utilized in the present invention, is 
substantially smaller than the width of a trunk type piston. The guide 
rail assembly of the present invention is less affected by temperature 
ranges because there is less metal to expand. 
In a preferred embodiment of the present invention, the guide rail is one 
twenty-forth (1/24th) as wide as a trunk type piston. Thus, the guide rail 
will expand twenty-four (24) times less than a trunk type piston having a 
diameter twenty-four times as wide as the guide rail and made of the same 
material, operating at the same temperature. Thus, an excess gap between 
the stabilizing member (the guide key) and the guide rail in which it 
resides could be twenty-four (24) times smaller than the gap between the 
conventional stabilizing member (the piston trunk) and the guiding 
cylinder wall in which it resides. This factor of twenty-four (24) gap 
tolerance clearance advantage manifests itself in the tolerance of the 
manufacturing process. The engine, of the present invention can be 
manufactured using reduced tolerance machining to enable manufacture of 
the engine to proceed quickly and without excess tolerance or induce wear 
caused by a loose fitting piston rattling in a typical engine. The piston 
of the present invention maintains a more consistent stability and 
decreases engine wear while enabling an overall smaller engine to be 
assembled. 
Turning now to FIG. 13A, the engine provides a piston crown 70 and cross 
head guide rail assembly 72. Piston thrust pads 74 are provided to center 
the crown in the cylinder. Connecting rod 76 engages connecting rod pin 
78, which abuts the bottom face of piston crown 70. Cross head rail guide 
72 is attached to piston crown 70 by bolts 80. As shown in FIG. 13C, guide 
keys 82, are provided in the lower crank case where guide keys 82 engage 
cross head guide rail slots 72. The cross head guide rail slots 72 and 
keys 82 stabilize piston crown face 86 and keep it perpendicular to the 
longitudinal axis of the cylinder parallel to the axis of translation of 
the piston crown within the cylinder. A plane drawn in the face of the 
piston crown 86, thus is kept perpendicular to the longitudinal axis of 
the cylinder. Thrust pads 74 maintain the piston crown 86 centered within 
the cylinder in which it resides. FIG. 13C is a sectional view of the rail 
guides 72 and keys 82. The slot guide passes along side the crankshaft 
enabling a shorter height cylinder and thus a shorter engine to be 
manufactured using the engine of the present invention. 
FIG. 14 shows the cross head guide rail assembly for the piston. The guide 
slot 72 of the cross head guide assembly engages the key 82. As shown in 
FIG. 14, the guide rail keys 82 protrude along the crank case wall along 
the longitudinal axis of the crankshaft 68. The center line 88 of the 
cylinder is shown for reference. 
The piston utilized in the present example of a preferred embodiment 
provides several advantages over typical trunk type pistons. As shown in 
FIGS. 15A and 15B, the typical connecting rod pin 96 is located a distance 
98 below the bottom of the piston 94. Typical connecting rods 100 are 
attached to connecting pin 96. In the typical trunk type piston, as shown 
in FIG. 15A, the force of combustion 92 presses down on the top of 
conventional trunk type piston 94. As shown in FIG. 15B, the combustive 
force 92 pressing down on conventional piston top 94 places a force and 
induces an associated bending moment on connecting pin 96. This bending 
moment tends to stress connecting pin 96, trying to bend connecting pin 96 
around the longitudinal axis 102 of connecting rod 100. This bending 
moment tends to place undue wear on the connecting pin and shortens engine 
life. There is no bending induced on the connecting pin of the piston 
assembly provided by the present invention. 
Turning now to FIG. 16A, in the piston of the present invention, force 92 
acts on the top of the piston crown 86. In the present invention 
connecting pin 78 adjoins both the lower surface of the piston crown 86 
and the top of connecting rod 76. Thus, there is no bending moment applied 
to the connecting pin 78, as it mechanically engages both the bottom of 
the piston crown 86 and connecting rod 76. The connecting rod 76 and 
connecting rod pin 78 are attached to piston crown 86 by retaining rings 
104. FIG. 16B is a view of the piston crown connected to the connecting 
rod and connecting pin turned ninety degrees (90.degree.) from the view 
shown in FIG. 16A. 
Connecting Rod 
Turning now to FIG. 17, in the present example of a preferred embodiment of 
the present invention, the engine provides a connecting rod 76. The 
smaller end 116 of connecting rod 76 attaches to the bottom surface of the 
piston crown as shown in FIG. 13A, discussed earlier. The large end of the 
connecting rod 107, as shown in FIG. 17, connects to the crank pin 108, as 
shown in FIGS. 18A and 18B. 
Turning now to FIGS. 19A and 19B, in the present example of a preferred 
embodiment, the large end of the connecting rod forms a 136.degree. 
semicircular arc which closely approximates the outside diameter of tabbed 
bearing 124. Tabbed bearing 124 abuts connecting rod end 107 on its 
outside diameter and the crank pin 108 on its inside diameter. The tabbed 
bearing 124 provides oil apertures 122 which enable oil to pass to provide 
lubrication for the connecting rod crank pin assembly. 
As shown in FIG. lSB, the width 126 of connecting rod end 107 is preferably 
a minimum distance to enable hydrodynamic bearings to be formed between 
the connecting rod end 107 and the tabbed bearing 124. The tabbed bearing 
also provides for lubrication between the internal diameter of the tabbed 
bearing 124 and the crankshaft pin 108. 
As shown in FIGS. 19A and 19B in the present example of a preferred 
embodiment, a plurality of connecting rods 76 are attached to one 
crankshaft pin 108. Connecting rods 107 preferably do not encircle 
crankshaft pin 108. Thus each connecting rod end 107 requires less than 
180.degree. of crank pin surface. In the present example, they are each at 
a 136.degree. arc. The connecting rod ends 107 may rotate relative to 
crankshaft pin 108 without interfering with each other. A set of retaining 
rings 126 (not illustrated for clarity) are utilized to rotationally 
attach connecting rods ends 107 around the tabbed bearing 124 and the 
crankshaft pin 108. The modular crankshaft pin comprises a male and female 
member which are inserted through the circular opening in the tabbed 
bearing 124 after the connecting rods 76 and retaining rings (FIG. 18A) 
have been assembled to form a circular structure around the connecting rod 
assembly. 
The connecting rod assembly of the present invention enables an engine 
designer to connect more than one connecting rod to a single crank pin. 
Multiple connecting rods can be attached to a single pin while utilizing a 
minimum length crank pin just long enough to accommodate lubricating a 
connecting rod of minimum width 126, as shown in FIG. 18B. The minimum 
crank pin length is preferably equal to the minimum width for which a 
single connecting rod 107 has adequate lubrication. The minimum width 
crank pin of the connecting rod assembly enables the engine designer to 
build a shorter crank pin and overall shorter crankshaft. Each crankshaft 
pin length in the crankshaft is reduced by a factor equal to the number of 
connecting rod ends attached to an individual crankshaft pin. A shortened 
crank pin reduces the bending moment asserted on the crank pin. The 
shorter crankshaft experiences smaller bending moments for a given force 
than a longer crankshaft. 
Turning now to FIG. 20, the crankshaft pin 108 is provided having large 
diameter crankshaft pins 109 and crankshaft 114 to reduce the torsional 
deflection induced in the crankshaft by the forces applied by the pistons. 
The crankshaft provided is a plurality of modules which plug together. 
After the connecting rod assembly has been assembled, male and female 
sections of the crank pin can be inserted and joined inside of the 
circular end of the connecting rod assembly. The connecting rod ends 107 
do not fully encircle the crankshaft pin 108 so that gaps 110, 112 and 
114, as shown in FIG. 18A, are formed between the connecting rod ends 107. 
Tabbed bearing 124 is utilized for lubrication between the crankshaft pin 
108 and the connecting rod end 107. Tab 120, on tabbed bearing 124, 
restricts the rotational motion of the tabbed bearing 124 and oil 
apertures 122 relative to the connecting rod ends 107. Thus, oil apertures 
122, which supply oil to the exterior surface of the tabbed bearing and 
the interior surfaces of the connecting rod end 107 are prevented from 
rotating far enough to become exposed to the gaps 110, 112, 114 between 
the connecting rod ends 107. Thus, oil is prevented from being pumped from 
the oil apertures 122 and through gaps 110, 112 and 114. Oil pumped 
through the gap flows to the bottom of the engine and has to be recovered 
with a scavenging pump. Reduction of the amount of oil escaping through 
the gaps reduces the amount of oil that has to be pumped back. The 
scavenging pump can be smaller in the present invention. This reduces the 
over all engine size. The positioning of the oil apertures 122, so that 
they stay under the connecting rod ends 107, and do not allow oil to 
escape through the gaps 110, 112 and 114. This also reduces the amount of 
oil which must be supplied to the connecting rod assembly by the supply 
pump. This reduces the size of the oil supply pump required to pump oil to 
the connecting rods and thereby reduces the overall size of the engine. 
In an alternative embodiment of the present invention, a male eyelet and 
female circular eyelets are formed in the bottom of connecting rods which 
share a crankshaft pin. The female circular eyelet comprises a forked set 
of circular eyelets which slide over the male circular eyelet. The 
combined male and female eyelets form a circular eyelet connecting rod 
assembly. The assembly is of a width sufficient to enable formation of 
hydrodynamic bearing between the crank pin and the connecting rod ends 
that slides over the crankshaft pin. 
In the alternative embodiment, a pressed sleeve bearing is press-fitted 
onto the forked female eyelet so that the pressed bearing sleeve does not 
rotate relative to the female connecting rod. The crankshaft connecting 
pin rotates underneath the sleeve bearing of the female rod. The 
displacement between the male connecting rod end and the sleeve bearing 
pressed into the female connecting rod end is thus minor. The male rod 
rotates over the bearing fixed in the female eyelet as the crankshaft 
rotates in a circle within the bearing fixed in the female eyelet. Two 
connecting rods drive two pistons by driving a single connection formed by 
the male and female connecting rod ends. 
The path of the connecting rod ends 107 and tabbed bearing 124 utilized in 
the present example of a preferred embodiment is illustrated in FIG. 19A 
and 19B. FIG. 19A shows the connecting rod ends 107 and tabbed bearing 124 
when the crankshaft has rotated to bottom of the piston stroke. At this 
point, the center of the crankshaft is at point 130, as shown in FIG. 19A. 
In FIG. 19B, the crankshaft has now rotated to the top of the piston 
stroke and the center of the crankshaft pin is now located at point 132. 
Point 130 is repeated for reference. Notice that in FIG. 19A, when the 
pistons are at the bottom of the stroke, the connecting rod ends 107 do 
not meet but leave a gap 110 between them. In FIG. 19B, when the pistons 
are at the top of their stroke, the connecting rod ends 107 rotate so that 
they leave small gaps 114 and 112 (FIG. 18A) between the connecting rod 
ends 107 and tab 120 of tabbed bearing 124. Oil apertures 122 remain 
underneath connecting rod ends 107 and are not exposed to gaps 110, 112 or 
114 during any point of the rotation of the crankshaft pin. 
Modular Crankshaft 
Turning now to FIG. 20, in the present example of a preferred embodiment, 
the engine utilizes a modular crankshaft 114, as shown in FIG. 20. The 
modular crankshaft 114 utilizes a male 111/female 109 assembly to form a 
crank pin 108. The male section 111 slides into the female section 109 to 
form crank pin 108. The male and female sections are splined together for 
rotational fixation between them. The present invention provides a 
structure which reduces the bending moment asserted on the crank pin 108 
by the connecting rod end 107. This is accomplished by reducing the width 
126 (FIG. 18B) of crank pin 108, to the minimum width needed to form a 
hydrodynamic bearing, based on the width of a single connecting rod end 
107, tabbed bearing 124, and crank pin 108. The necessary length of the 
crank pin is reduced because more than one connecting rod end 17 is 
attached to the pin 108. 
Two connecting rod ends 107 are connected to a single width 126 crank pin 
18, reducing the necessary overall length of crank pins by a factor of 
two, because two connecting rod ends are sharing the same crank pin whose 
length equals the minimum width 126 of a single crankshaft pin. If three 
connecting rod ends 107 are connected to a single crank pin 108, the pin 
length requirement is reduced by a factor of three, and so on. 
As shown in FIG. 21, reducing the crank pin length reduces the overall 
crankshaft length and thus reduces the bending moment asserted across the 
width of a crank pin by connecting rod 107. Reducing the bending moments 
by minimizing the width utilizing a single width crank pin for multiple 
connecting rod ends, reduces the length of crank pins and thus reduces the 
overall length of the crankshaft. The reduced width of the crank pins 
reduces the bending moment of a force asserted on a crank pin. Thus, the 
crank pins suffer less deformation twisting, and torsional deflections 
during operation. Crank pin 108 and crankshaft 114 are formed of large 
diameter tubing which minimizes the torsional deflection within the 
crankshaft and crank pins. 
The present invention provides sufficient overlap between the male section 
111 of the crank pin and the female section 109 of the crank pin. The 
crankshaft is made of a stiff material and is configured in large diameter 
so that the natural frequency of vibration of the crankshaft and crank 
pins is much higher than the frequency of rotational power impulses 
applied to the crankshaft by the low RPM pistons through the connecting 
rods. The present example of a preferred embodiment, utilizes a modular 
engine with a maximum RPM of approximately 2,700. Thus, the frequency of 
piston impulses applied to the crankshaft is much lower in the low RPM 
engine than the natural frequency of vibration of the large diameter 
crankshaft. The frequency of the impulses supplied by the pistons does not 
match the natural frequency vibration of the crankshaft of the present 
invention. This mismatch substantially reduces the possibility of harmonic 
breakage of the crankshaft to lower than that encountered with typical 
modular crankshafts. 
As shown in FIG. 21, crankshaft bearings 134 and 136 support each section 
115 of the modular crankshaft. Each section of the modular crankshaft is 
support by bearings 134 and 136, so that the bending moments and shear 
forces from the piston are resolved in a redundant manner by each of the 
crankshaft sections 114 and 115, which connect together to form a crank 
pin 108, which receives the load from a piston. 
Cylinder Head Seal 
Turning now to FIG. 22, in a preferred embodiment of the present invention, 
the engine utilizes a metal to metal seal between cylinder head 138 and 
cylinder 141. Cylinder head 138 is configured separately from cylinder 
141. The shell of configuration of the separate cylinder head 138 enables 
conventional machine bits to traverse the depth of the cylinder head 138. 
The shallow depth of the cylinder head 138 enables short rigid machine 
bits to accurately machine the cylinder head surfaces. Longer machine 
bits, which would be required with a one-piece cylinder head and cylinder 
would have to traverse the length of the cylinder to reach and machine the 
cylinder head. Configuration would require long machine bits which would 
be less rigid and thus less accurate in machining of the cylinder head 
138. 
The cylinder head of the present invention utilizes a metal to metal seal 
between chamfered edge 146 of cylinder 141 and a flat surface 150 within 
the female portion of cylinder head 138 into which the top male portion of 
the cylinder inserts. The present invention has advantages over cylinders 
assembled using gaskets to seal the cylinder head. The metal to metal 
contacts of the present invention forms a seal without the attendant 
variations in assembled tolerances experienced when utilizing gaskets to 
assemble an engine. 
The cylinder head 138 forms a female receptacle into which the wall of 
cylinder 141 slides and mechanically engages. Chamfered edge 146 of 
cylinder 141 abuts flat surface 150 of the cylinder head 138. Turning now 
to FIG. 23A, retaining ring 142 is shown as a U-shaped bracket, forming 
right angles 144 and 143, and fitting over cylinder head land 139 and 
cylinder wall land 140. Retaining ring ends 143 and 144 abut lands 139 and 
140. 
Turning now to FIG. 23B, an indention 145 is then formed in retaining ring 
142. This indention 145 shortens the retaining ring 142 so that retaining 
ends 143 and 144 are drawn closer together. Retaining ring 144 exerts a 
compressive force on cylinder head land 139 and cylinder wall land 140 
bringing the two lands closer together and applying a compressive force on 
cylinder chamfered edge 146 which opposes the flat surface 150 of the 
cylinder head. The pressure asserted by the retaining ring on the 
chamfered edge 146 forms a seal between the chamfered edge 146 and the 
flat surface of the cylinder head 150. Thus, a metal to metal seal is 
formed in the combustion chamber between the cylinder edge 146 and 
cylinder head surface 150. The retaining ring and lands form a flange 
which fits into a female groove 195 formed in engine half 26 (FIG. 5). 
The combustion pressure between the top of the piston crown and the 
interior surface of the combustion chamber formed by the cylinder wall and 
the interior of the cylinder head tends to assert a force on the cavity 
formed between the chamfered edge 146 and flat surface 150. Pressure 
within this small area is negligible and not threatening to the integrity 
of the seal between the cylinder and cylinder head. Any combustive force 
that leaks through the cylinder head seal, if any, exerts a negligible 
pressure on the gap formed between chamfered edge 146 and flat surface 
150. 
Lubrication System 
In the present example of a preferred embodiment of the present invention, 
the engine provides an independent lubrication system for each engine. 
Referring to FIG. 4, each engine contains an independent lubrication and 
cooling system comprising a coolant pump 402, a scavenger pump 404, and a 
pressure pump 406. When engines are interconnected, the coolant and 
lubrication fluids are manifolded 408a, 408b and 408c in parallel to each 
engine so that each engine is supplied with lubricating and coolant fluid 
at the same temperature. Thus, each engine runs at the same temperature. A 
plurality of engines connected together to form an extended engine, will 
have an adequate pumping system because each engine is independently 
lubricated and cooled. There is no need to add additional pumps to an 
assembly of interconnected engines other than to manifold the supply to 
the engines. 
The present invention has an advantage over typical engines which supply 
coolant and lubricant serially to each engine. Typical engine designs 
provide for serial coolant and lubricate distribution. Coolant and 
lubricate are first run through a first engine before they are run through 
the second, third, fourth, fifth engine, etc. In the present example of a 
preferred embodiment, the coolant and lubricant are provided in parallel 
to each engine module so that the coolant and lubricant are supplied to 
each engine at the same temperature, rather than preheating the lubricate 
and coolant in the first engine before sending it to the second engine and 
so on. Thus, the present invention runs at a lower overall temperature and 
a more constant temperature. The inherent adequacy of the independent 
pumps provided within each engine eliminates the need for an engine 
manufacturer to install custom pumping systems to promulgate various 
numbers of engines connected together to form a engine. 
Valves 
As shown in FIG. 24, in the present example of a preferred embodiment of 
the present invention, three intake valves 152 and three exhaust valves 
153 are provided per cylinder head 154. Spark plugs 156 are shown in FIG. 
24 for reference. The use of six valves, combined with the low RPM of the 
engine enables the cylinder head of the present invention to perform using 
a very small opening under the valve. As valves are lifted only a short 
distance and are elliptical or flattened ports to induce tangential gas 
flow. Six valves generate a large contact area relative to the overall 
valve mass and area. Thus, the design of the valves in the present 
invention enable rapid heat transfer from the valve to the head. 
Turning now to FIGS. 25 and 26, the configuration of the ports is such that 
all surfaces may be machined with standard milling bits operating from an 
axis parallel to the valve axis and parallel to the port axis. The shape 
and angle of the valve housing is such that small valve openings enable 
gas flow which reacts to the valve more as a streamline rather than as a 
90.degree. impediment. The entire head above the combustion chamber is 
pressurized with cooling oil, thus the valve stem spring and cam follower 
mechanism are immersed in coolant. The flattened or elliptical ports allow 
for a short heat path to the coolant, as shown in FIGS. 25 and 26. 
FIG. 27 shows the crank case engine half, guide key 82. While an example of 
a preferred embodiment of the present invention has been presented, it is 
not in tended to limit the spirit or scope of the invention. Variations of 
the preferred embodiment are possible while remaining within the scope of 
the claimed invention.