Gas turbine-two-stroke piston compound engine

A compound engine including a gas turbine unit and a two-stroke uniflow scavenge piston unit having a hollow cylindrical exhaust valve mounted concentrically with an associated piston to form a portion of the combustion chamber within a peripheral wall of the exhaust valve. A double acting exhaust valve actuator is responsive to rotation of the crankshaft and is operatively coupled to the peripheral wall of the hollow cylindrical exhaust valve to produce forces derived directly from rotation of the crankshaft to open and close the exhaust valve. The actuator is a camshaft having two sets of cams which drive a cam follower in response to rotation of the crankshaft.

FIELD OF THE INVENTION 
The invention relates generally to compound engines, and more particularly, 
to a single engine comprised of a combination of gas turbine and piston 
units which are drivingly connected to an output drive shaft. 
BACKGROUND OF THE INVENTION 
An early compound engine that is comprised of a gas turbine unit and a 
piston unit is disclosed in the Johnston U.S. Pat. No. 3,498,053. In 
Johnston, the turbine unit is operated by the exhaust gases from a 
reciprocating two cycle, compression ignition, piston unit. The gas 
turbine unit of Johnston has a compressor component to supercharge the 
piston unit, and the turbine unit is also coupled to the output shaft of 
the piston unit. The Johnston piston unit has pairs of opposed cylinders 
with a scotch yoke connecting the pistons to a crankshaft. 
A subsequent version of the compound engine is described in a publication 
entitled "A New Concept For Reduced Fuel Consumption In Internal 
Combustion Engines" presented at the 1971 Intersociety Energy Conversion 
Engineering Conference Proceedings, P38, by the Society of Automotive 
Engineers. This publication discloses a low pressure turbine mechanically 
coupled to the output shah of the piston unit and fluidly driven by the 
output of a high pressure turbine unit. The piston unit is a compression 
ignition two-stroke unit in which excess scavenge air flow in the 
valve-open portion of the stroke is used to internally cool the piston and 
cylinder. The exhaust valve is an annular valve located at the top of the 
cylinder and forms a portion of the wall of the combustion chamber. A 
bypass burner, that is, a thermal reactor, is placed in a parallel flow 
path between the high pressure compressor and high pressure turbine. To 
start the engine, a small starter motor spins the high pressure turbine, 
and combustion is initiated in the bypass burner. The bypass burner then 
becomes the combustion system to provide a driving gas for the high 
pressure turbine. The high pressure turbine first, mechanically drives the 
compressor to provide high pressure air to the piston unit, and second, 
provides a driving gas to the low pressure turbine which, in turn, 
provides cranking power to start the piston unit. 
While the feasibility of many of the principles of the above engine 
configuration has been proven, there are several areas where components 
that are suitable for either only turbine engines, or only piston engines, 
are less suitable for compound engines and thus require further 
investigation and development. For example, the exhaust valves in typical 
compression ignition engines are generally of the poppet type and are 
actuated by an engine driven cam shaft which operates the valves through a 
rocker arm arrangement. The valves are held in a closed position by means 
of compression springs against which the rocker arm and cam operate to 
open the valves. As engine speed increases, the rate of occurrence of 
combustion cycles increases; and the speed with which the exhaust valves 
are opened and closed also increases. The speed with which the exhaust 
valves open is controlled by the engine driven cam. However, the speed 
with which the exhaust valves close can only be increased by increasing, 
or making larger, the spring constant of the valve return springs. While 
the stronger valve return springs facilitate operation of the engine at 
high speeds, these very strong valve compression springs have the 
disadvantage of requiring a very high torque to rotate the engine for 
starting, thereby making the engine more difficult to start, or otherwise 
turn over with an auxiliary starting motor or the low pressure turbine. 
Therefore, to reduce the starting torque of the engine requires that the 
valve return springs be weaker; however, those weaker springs compromise 
the maximum speed of the engine because, as the engine speed increases, a 
speed will be reached at which the valve return springs will not be able 
to bring the valve to its fully closed position prior to combustion. 
Consequently, the exhaust valves will float in a constantly open position 
which has the disadvantage of governing or limiting the speed of the 
engine. 
Conventional compression ignition engines address the high starting torque 
problem in several ways. First, a starter motor powered by a battery is 
used to start the engines, and an increased starting torque is 
accommodated by simply increasing the size of the starting motor and 
battery. In addition, an igniter can be used to provide ignition while the 
cold engine is being started. Often a bleeder valve is operable when the 
engine is being started to bleed pressure from the cylinder during the 
compression stroke. While effective, all of the above solutions have the 
disadvantage of adding complexity, weight and cost to the engine. More 
importantly, such solutions interfere, or are incompatible with, a turbine 
and piston unit working together. 
Compound engines, and in particular the compound engines described above, 
have a piston unit with specially configured exhaust valves and combustion 
chambers. Further, the combination of the gas turbine unit with the piston 
unit presents difficult gas flow and energy transfer requirements. The 
valve actuations proposed by the prior art create additional problems in 
the manufacture and operation of such a compound engine. Accordingly, 
there remains a need for a solution to the valve actuation problems 
associated with compound engines of the type to which the present 
invention relates. 
SUMMARY OF THE INVENTION 
An object of the present invention is to provide a mechanism for operating 
the exhaust valves of the piston unit of the compound engine which 
eliminates the above compromise between engine speed and starting torque. 
A further object of the present invention is to provide a mechanism for 
operating the exhaust valve of the piston unit of the compound engine 
which allows the piston unit to operate over its full speed range. 
Another object of the present invention is to provide a mechanism for 
operating the exhaust valve of the piston unit of the compound engine 
which presents sufficiently low starting torque that a low pressure 
turbine of the gas turbine unit, which is connected to the piston unit 
crankshaft, can be used to start the piston unit of the engine. 
To overcome the above disadvantages, the present invention provides a 
compound engine comprised of a gas turbine unit and a two-stroke uniflow 
scavenge piston unit in which the piston unit may be operated over a wide 
speed range but may be started, or otherwise turned with the gas turbine 
unit. 
According to the principles of the present invention, there is provided a 
compound engine including a gas turbine unit and a two-stroke uniflow 
scavenge piston unit in combination with a double acting cam drive for 
operating the piston unit exhaust valves. In accordance with the described 
embodiments of the invention, a cam shaft is coupled to the crankshaft of 
the piston unit and contains two cam surfaces for each exhaust valve. A 
first cam surface is mechanically coupled to the exhaust valve and is 
operative to move the exhaust valve to an open position. A second cam 
surface is also mechanically coupled to the exhaust valve and is operative 
to return the exhaust valve to its closed position. 
In another embodiment of the invention, a first cam follower is in contact 
with the first cam surface to move the exhaust valve in a first direction, 
and a second cam follower is in contact with the second cam surface to 
move the valve in a second direction. In a further embodiment, the second 
cam surface includes a pair of cam surfaces on each side of the first cam 
surface; and the second cam follower includes a pair of second cam 
followers in contact with the pair of cam surfaces. In a further 
embodiment, the cam followers are mechanically coupled to the exhaust 
valve by a connecting link. In a still further embodiment, the exhaust 
valve is a cylindrical hollow body and the connecting link is a pair of 
connecting links connected to opposite sides of the exhaust valve. In a 
further embodiment, the invention provides a method of starting the 
compound engine with the double acting cam drive for the exhaust valve. 
And, in another embodiment of the invention, the double acting cam drive 
for the exhaust valve is applied to a two-stroke uniflow scavenge engine. 
The above double acting cam arrangement provides positive and direct 
exhaust valve operation that is synchronized with the operating speed of 
the piston unit. With the double acting cam, the exhaust valve will always 
be brought to its fully closed position and will not float at higher 
operating speeds. Therefore, the present invention has the advantage of 
not limiting engine speed. Further, by eliminating the valve return 
springs, the present invention has the further advantage of requiring less 
torque to start the piston unit; and the piston unit may be easily started 
with the gas turbine unit. 
These and other objects and advantages of the present invention will become 
more readily apparent during the following detailed description, together 
with the drawings herein.

DETAILED DESCRIPTION OF THE INVENTION 
Referring to FIGS. 1, 2 and 3, the compound engine 10 is comprised of a 
piston unit 12 and a gas turbine unit 36. The piston unit 12 is preferably 
a compression ignition, two-stroke, uniflow scavenge unit which includes 
opposed pairs of cylinders 14, preferably eight cylinders. The opposed 
pairs of cylinders 14 are arranged in two banks 16, one cylinder 14 of one 
bank 16 is directly opposite one cylinder 14 of the other bank 16. Each 
pair of the cylinders 14 is drivably connected to the crankshaft 18 by 
means of a scotch yoke 20. Each cylinder 14 contains a piston 22 rigidly 
connected to one end of the scotch yoke 20. The scotch yoke 20 has a 
crosshead 24 with a rectangular slot 26 that has a slider block 28 
slidably mounted therein which is rotatably coupled to an eccentric 30 of 
the crankshaft 18. The cylinders 14 are generally identical, and each pair 
of cylinders 14 with interconnecting scotch yoke 20 are generally 
identical. Similarly, the bank 16 of cylinders 14 generally identical, 
being mirror images of each other. The piston unit 12 of the compound 
engine 10 is a two-stroke compression ignition diesel unit to each 
cylinder 14 of which combustion or cycle air is fed through an intake 
manifold 32 and air intake ports 34 when its respective piston 22 is in 
the bottom-most portion of its stroke. 
The cycle air is supplied by a high pressure gas turbine unit 36 comprised 
of a steady flow high pressure compressor 38, a high pressure turbine 40, 
a pair of combustors 42 and an axial flow low pressure turbine 44. In the 
preferred embodiment, the high pressure compressor 38 receives cycle air 
through an inlet 46; and the air passes through vanes of a compressor 
rotor 47 and through a discharge scroll 48 that divides the compressed air 
into two discharge paths 49,49, each of which routes the air to one of the 
two combustors 42. The turbine unit 36 is configured such that exhaust 
gases from the cylinders 14 of each bank 16 of the piston unit 12 pass 
through one of a pair of exhaust manifolds 50, respectively associated 
with each bank 16, and through a respective one of the two bypass 
combustors 42 of the gas turbine unit 36. The combustors 42 are configured 
to drive the high pressure turbine 40 by routing the exhaust gases from 
the combustors 42 to the two entrances on each side of the engine of a 
dual inlet variable area turbine nozzle scroll 52 and through the vanes of 
the high pressure turbine rotor 53. The high pressure turbine 40 output 
shaft is connected to a bearing and shaft assembly 54 to drivably rotate 
the high pressure compressor 38. The high pressure turbine 40 has variable 
inlet vanes (not shown) which vary the inlet area of the high pressure 
turbine to permit the high pressure turbine 40 to run at 100% of its 
desired speed even at lower speeds of the piston unit 12 which produce 
significantly less exhaust gases. The outlet of the high pressure turbine 
40 is fluidly connected to a conically shaped diffuser 56 which is 
connected to the inlet of the axial flow low pressure turbine 44. Variable 
vanes (not shown) may be utilized on the input of the low pressure turbine 
44 to vary the pressure drop and, hence, the power produced by the low 
pressure turbine. Output gases from the low pressure turbine 44 are vented 
to atmosphere through the outlet 58. 
The low pressure turbine 44 has a unitary rotor 59 and output shaft 60 
connected through a speed reduction unit 62 the output of which is 
mechanically coupled to the crankshaft 18. The speed reduction unit 62 
mounts directly on the low pressure turbine unit 44 and is lubricated by a 
common system. The speed reduction unit 62 consists of a ring gear, three 
planet gears and a sun gear to provide a very low friction, 10:1 speed 
reduction on the same shaft centerline as the low pressure turbine 44. A 
2:1 speed reduction is accomplished in the pulley and V-belt arrangement 
64 connecting the output of the speed reduction unit 62 with the 
crankshaft 18. A flywheel 66 weighing approximately 120 pounds is also 
mounted on the crankshaft 18 which provides rotary shaft output power from 
the compound engine 10. 
The scotch yoke 20 is rigidly connected to the pistons 22 and the centrally 
located rectangular slot 26 extends longitudinally in a direction 
perpendicular to the stroke of the opposed pistons 22. That arrangement 
has several advantages. First, the Scotch yoke 20 provides positive 
guidance for the pistons 22 as they reciprocate in the cylinders 14, 
thereby minimizing the lateral forces between the pistons 22 and the 
cylinders 14. Therefore, friction and wear between the piston rings (not 
shown) and cylinders 14 is minimal. Second, for uncooled operation, the 
top of the piston and cylinder can approach 1700.degree. F. With the rigid 
scotch yoke 20, the pistons 22 are guided within the cylinder 14 by 
clearance control on the cool and lubricated base of the piston. With the 
scotch yoke 20, piston skirts can be very short and provide guidance for 
the piston 22. A clearance between the piston and cylinder above the 
skirt, for example 0.005 inches, can be maintained causing no additional 
rubbing of the hot piston and cylinder parts. Third, the expansion forces 
of combustion of one piston is transferred directly as a compression force 
to the opposing piston. In addition, with an eight cylinder unit, two of 
the scotch yokes 20 are always moving at the same speed, but in opposite 
directions, which results in a balanced reciprocating unit. Preferably the 
scotch yoke is made from aluminum or composite material and has a total 
reciprocating mass of approximately 14 pounds or less. 
The maximum bearing load occurs when the slider block 28 is in the center 
of its run and its velocity is maximum. A hydrodynamic bearing film at 
this condition is approximately 4000 pounds per square inch ("psi"). At 
each end of the slider run where it reverses direction, the hydrodynamic 
film disappears and must be reestablished as the slider starts its return. 
To enhance the reestablishment of the hydrodynamic lubrication film, 
referring to FIG. 4, a pivoting pad or skate 68 is pivotally mounted on 
each side of the slider block 28. Each pad 68 is pivotally supported by a 
trunion, or pin 70, and each pad 68 bears against a metal insert or spring 
72 mounted in the slider block 28. The tilting pad or skate 68 is 
compressed against the bearing surfaces within the slot 26 of the 
crosshead 24 with a force of 400 pounds by deflection of the metal insert 
or spring 72. During operation of the piston unit 12, as the skate 68 
reverses direction, oil is ported through drilled holes (not shown) in the 
crankshaft 18 and slider 28 and applies a force to the trailing end of the 
skate 68, for example, the left end as viewed in FIG. 4, causing the skate 
to tilt slightly and open a wedge 74 on its forward edge, thereby rapidly 
reestablishing the hydrodynamic oil film. At the center of each slide, the 
film is determined to be approximately 0.0008 inches on the leading edge 
and 0.0004 inches on leaving the skate 68. 
To summarize the operating cycle, referring to FIGS. 1 and, 2, with the 
above compound engine, high volume, low pressure air is compressed by a 
total ratio of approximately 200:1. The cycle air is first compressed by a 
ratio of approximately 5:1 by the rotating high pressure compressor 38 
after which air flows through the combustors 42, the intake manifold 32, 
intake ports 34 and into the cylinders 14 of the piston units 12. The air 
is further compressed by a ratio of approximately 40:1 when the piston 
units 12 are operating at full power. The compression ignites fuel 
injected into the cylinders 14 near the top dead center portion of the 
piston cycle, and the energy of the combusting and expanding gases is 
extracted to the maximum extent possible at nearly one hundred percent 
(100%) efficiency by the piston units 12 through a crankshaft rotation of 
approximately 95.degree. past top-dead center and an additional 30.degree. 
during the opening of the exhaust valve 110. When the gases have been 
fully expanded in the cylinders 14 and combined with the cooling and 
scavenge air, they are returned through the combustors 42, to drive the 
high pressure turbine 40 which in turn rotates the high pressure 
compressor 38. Energy remaining in the exhaust gases from the piston units 
12 is extracted in the low pressure turbine 44 which is connected through 
the gear reduction unit 62 and a V-belt unit 64 to the output of the 
crankshaft 18. 
The two-stroke cycle of the piston unit 12 is important because the cycle 
air flow characteristics of the two-stroke cycle more nearly match the 
continuous gas flow characteristics of the gas turbine unit 36. For 
instance, the intake and exhaust valves of the piston unit 12 are open for 
large portions, for example, one-third, of the rotation of the crankshaft 
81. The cycle air flow within the compound engine 10 is unique because of 
the parallel flow paths provided by the combustors 42. As shown in FIG. 5, 
cycle air from the high pressure compressor 38 enters an outer first inlet 
80 intersecting a cylindrical flow path 81 at the rear end of the 
combustors 42 and flows to its forward end. The cycle air then passes 
through annular port 82 and into a second cylindrical flow path 84. The 
second cylindrical flow path 84 provides a first air outlet that 
intersects the intake manifold 32 and radially extending struts or air 
passages 86 of a circumferential valve 88. The circumferential valve 88 
includes a sleeve 89 which is slidably located over the struts 86 and has 
openings matching the openings of the strut air passages 86. During normal 
engine operation the sleeve 89 of the circumferential valve 88 is rotated 
to a position that closes but does not seal the air passages 86, and the 
air flows from the annular passage 81, through the intake manifold 32 and 
into ports 34. The sleeve 89 of the circumferential valve 88 is rotated to 
open the strut air passages 86 when operating bypass burners, or heaters, 
within the combustors 42 of the gas turbine 36. With the combustor 
operation, the air passages 86 intersect in an inner cylindrical air flow 
channel 90 which by means of cycle air ports 92 provide combustion air to 
a combustion chamber 94 comprising the bypass burner, or heater, within 
the combustors 42. Fuel is injected into the combustion chamber 94 by 
means of an injection line 96. Burning fuel in the thermal reactors, or 
combustors 42 produces exhaust gases to the second outlet passage 106 
which operate as driving gases for the high pressure turbine 40. The 
bypass burners in the combustors 42 are ignited when starting the piston 
unit 12 or when it is desirable to provide a power boost to the compound 
engine 10. A unique feature of the compound engine 10 is the parallel air 
flow paths within the combustors 42. For example, cycle air can flow both 
to the piston unit 12 and to the combustion chambers 94 of the combustors 
42. By simultaneously operating the piston unit 12 and burning fuel in the 
bypass burners or combustion chambers 94 of the turbine unit 36, both 
units 12, 36 supply driving gases to the high pressure turbine 40 which in 
pass through outlet 43 (FIG. 2) and turn, provide a substantial increase 
in the driving gases to the input 45 of the low pressure turbine 44. The 
output drive shaft 60 from the low pressure turbine 44 adds significant 
power to the crankshaft 18 from which the output power from the compound 
engine 10 is taken. 
In another flow path, exhaust cycle gas from combustion within the 
cylinders 14 exits cylinders 14 via exhaust manifold 50 and enters a 
second inlet of the combustors at 98. The exhaust cycle gas, from the 
piston unit 12 flows through longitudinal passages 100 that extend between 
the struts or radial air passages 86. At the other end of the combustors 
42, the exhaust cycle gas and the intake cycle air flowing through the 
cylindrical longitudinal air passage 102 pass through a daisy mixer 104 
which breaks the exhaust cycle, gas and intake cycle air into smaller 
streams so that they more quickly mix and merge into outlet passage 106 
that is connected to the inlet of the high pressure turbine 40. 
Air from the high pressure compressor 38 flows through the combustors 42 to 
both the intakes of the open cylinders on the piston unit 12 and to the 
high pressure turbine 40. This operation provides compressed and heated 
air to the intake manifold 32 and torque to the crankshaft 18 through the 
low pressure turbine 44 which is fluidly coupled to the high pressure 
turbine 40. A unique aspect of the parallel flow path provided through the 
combustors 42 is that the gas turbine 36 may be started and operated 
independently of the piston unit 12 thereby providing the advantages of 
easy starting and internal cooling flow path control. 
The unique air flow configuration is possible because both the gas turbine 
unit and the two-stroke, uniflow scavenge piston unit have a total 
pressure drop across the system to enable operation. Typical gas turbines 
have a pressure drop between the compressor and turbine of approximately 
five percent (5%). Recuperated turbines are higher. The pressure drop 
across a typical two-stroke cylinder varies with the valving arrangement, 
speed and power setting. For an unit with a fixed displacement scavenge 
compressor, the pressure ratio can vary from very low values at idle to 
perhaps forty percent (40%) at full power and full speed. Preferably, for 
surface applications of the compound engine 10, an approximately 5:1 
pressure ratio high pressure compressor 38 with an approximately 
eighty-two percent (82%) peak efficiency is preferred. The compressor 
efficiency is an important parameter for a gas turbine and diesel compound 
engine. Since the compressor 38 provides air to the piston unit at about 
400.degree. F., a lower pressure ratio will reduce the exhaust energy 
recovered in the high pressure turbine 40. On the other hand, a higher 
pressure ratio requires the piston compression ratio to be lowered to 
maintain reasonable peak cylinder pressure. In addition, as the compressor 
ratio increases, the air temperature furnished to the piston unit 12 
increases thereby reducing the cooling capabilities of that air. Further, 
the temperature of the cycle air at the intake manifold has a large effect 
on the volumetric efficiency, or the ability of the cylinder to obtain a 
sufficient charge of air on each stroke. 
With the pressure ratio across the cylinder of the piston unit 12 at about 
five percent (5%), and piston air flow varying as a function of the 
pressure drop and speed of the piston unit 12, variable area nozzles are 
used on both the high pressure turbine 40 and low pressure turbine 44. 
Therefore, the high pressure rotor operates at one hundred percent (100%) 
or full speed from full power down to about twenty percent (20%) power. 
With the high pressure rotor operating at full speed over most of the 
power range, high pressure air is furnished to the intake manifold 32; and 
a five percent (5%) pressure drop is maintained across the cylinder 14 at 
most power settings. With this arrangement, the scavenge ratio will vary 
from about five hundred percent (500%) near idle to one hundred twenty 
percent (120%) at full power and full speed. The scavenge ratio or 
trapping efficiency is highly influenced by the speed of the piston unit 
12. Scavenge efficiency is also influenced by the scavenge ratio. 
Another advantage of the air flow configuration is that it provides a 
convenient method of starting even at very low ambient temperatures. The 
high pressure turbine 40 is easily started by a small starter motor (not 
shown) to bring it up to, for example, 5% of its desired operating speed. 
At the same time, the circumferential valves 88 are opened to port cycle 
air through the radial struts 86 (FIG. 5) and igniting the combustors 42. 
Once the combustors 42 are burning, they supply air to rotate the high 
pressure turbine 40; and the high pressure turbine 40 can be disengaged 
from the starter motor. The high pressure turbine 40 supplies air to 
rotate the low pressure turbine 44 which is coupled to the crankshaft 18 
and provides a high torque which cranks the piston unit 12. The high 
pressure turbine 40 is also rotating the high pressure compressor 38 which 
by means of working the inlet air is supplying warm air to the intake 
manifold 32. The low pressure turbine 44 can crank the piston unit 12 for 
extended periods at high speeds until combustion ignition of the diesel 
cycle is achieved. At that point, the combustors 42 are shut off; and the 
circumferential valve 88 is closed. 
One of the design objectives of the compound engine 10 is to minimize the 
engine cooling and optimize the use of heat which is created during the 
compression combustion process. Therefore, the internal piston and 
cylinder surfaces preferably tolerate the 1500.degree. F.-1700.degree. F. 
surface temperatures. The efficiency and performance of the compound 
engine 10 is improved by its high temperature operation. Further, the heat 
flow characteristics of the piston unit are improved because the 
relatively cool compressor discharge air enters the bottom of the 
cylinders 14 and the exhaust cycle gas leaves the cylinders at their top 
end (horizontally displaced from the inlet). This arrangement makes it 
easier to maintain a relatively cool area in the piston unit crankcase 108 
of FIG. 3 to minimize heat transfer to the lubricating oil. It also 
provides an even circular temperature gradient in the cylinders thereby 
preventing deformation and distortion. Supplemental cooling is provided by 
the lubrication oil in the crankcase 108. A series of oil spray nozzles 
(not shown) are situated within the crankcase 108 to impinge oil on all 
scotch yoke bearing surfaces and on the underside of each piston 22. The 
crankshaft 18 is drilled to provide pressure lubrication to each main 
bearing and the internal slider bearings. 
Because of the requirement for minimal internal cooling, a low pressure 
drop through the cylinder 14, a very high peak cylinder pressure, and hot 
metal temperatures, the compound engine 10 has several unique design 
features. First, the opposing cylinders 14 are exactly 180.degree. apart 
and on the same centerline, and the cylinders are tied together with 
high-strength through bolts extending between the opposed cylinder heads. 
This provides for easy assembly and disassembly and relieves the 
longitudinal stress in the cylinder walls. This arrangement further 
provides larger intake port areas. The long, high-strength tie rods also 
permit the cylinder liner and head to be stacked in sections which permits 
for the easy insertion of insulating or low heat transfer gaskets, such as 
zirconia, if desired, to minimize the transfer of heat down the cylinder 
liner walls. 
Second, the piston unit 12 is designed as a uniflow scavenge unit wherein 
the cylinder 14 and piston 22 are tapered toward the top, thereby reducing 
the internal volume of the combustion chamber at its upper end in order to 
provide several advantages. With the location of the intake ports 34 at 
the bottom of the cylinders and the exhaust valves 110 at the top of the 
cylinders, the design provides an initial swirl of the cycle air at the 
intake ports. The swirling pattern of the intake air continues as it rises 
through the cylinder 14 and accelerates as it is squeezed to a smaller and 
smaller diameter as it moves up the conical cylinder volume. The 
combustion chamber takes the shape of a small cylindrical plug with 
reduced surface to volume area ratio for a given clearance volume. These 
factors along with the high temperatures of the combustion chamber 
surfaces provide for a high heat release configuration. Further, the rate 
of heat release from the surfaces within the cylinder are greatest at 
those areas where the temperature is highest. In addition, the reduced 
volume at the upper end of the cylinder facilitates the compression 
ignition process. Advantageously, ignition delay is eliminated with 
operating surface temperatures over 1000.degree. F. 
The reduced diameter combustion chamber also facilitates the use of a 
hollow, cylindrical exhaust valve 110 mounted within the cylinder head 112 
concentrically with the piston 22 (FIG. 3). With such a valve, the entire 
360.degree. circumference at the top of the cylinder is opened to provide 
an aerodynamically-shaped exhaust scroll providing a very low pressure 
drop across the cylinder. When the valve is closed, the exhaust valve 110 
bears against and seats on an annular surface 113 within the cylinder head 
112; and the combustion chamber is located inside the cylinder and the 
exhaust valve. The exhaust valve 110 also has an inside lip 114 on one end 
of its cylindrical annular body 158 which is oriented at an angle of 
approximately 30.degree. F. with respect to the horizontal and is used to 
provide a positive seating force during combustion when there is maximum 
pressure within the cylinder. An advantage of such a valve design is that 
scavenge efficiency is minimally from seventy percent to ninety percent. 
Further, depending on a combination of gas turbine and piston unit speed, 
the scavenge efficiency will reach one hundred percent (100%). 
The exhaust valve 110 is mounted around and moves longitudinally with 
respect to a center body 116 which holds a fuel injector 118. Fuel 
injection is accomplished by utilizing an eight-plunger fuel pump (not 
shown) with cam plunger springs and governor to drive the eight fuel 
injectors. All eight high pressure fuel injection lines are identical in 
length so that all injector needle lift pressures are approximately the 
same, for example, 3200 psi. Sealing rings 120 are contained on both the 
center body and the cylinder head 112 to seal combustion gases from 
leaking past the exhaust valve 110 that is reciprocating therebetween. By 
mechanisms not shown, the center body 116 may be selectively moved a small 
displacement toward and away from the piston 22, thereby changing the size 
of the combustion chamber within the piston unit 12 and hence the 
compression ratio. Changing the compression ratio is beneficial when 
operating the piston unit 12 at very low speeds, for example, at idle. 
With the designed pressure drop across the two-stroke piston unit, all 
cooling is provided by the cycle air and except for the scotch yoke 
lubrication cooling, no external cooling mechanisms are used. Small 
amounts of cycle air are diverted through passages 122 of the cylinder 14, 
passages 124 of cylinder head 112 and passages 126 of center body 116 to 
cool critical areas the piston unit 12. The cooling passages 124 are an 
annular arrangement of impingement cooling holes to direct air onto the 
tip of the exhaust valve 110 and the exhaust valve seat. The cycle air 
cooling is effective to maintain component temperatures at desired levels. 
Temperature levels are maintained high, for example, 1200.degree. 
F.-1500.degree. F. which reduces the amount of bypass cooling air and 
allows it to pick up sufficient energy such that upon reentering the cycle 
in the exhaust manifold much of the cooling energy can be recovered in the 
downstream turbine 44. Using that diverted cooling cycle air, the tip of 
the exhaust valve 110 is cooled by impingement air from the cylinder 
cooling holes 124. These holes are arranged so that they flow and impinge 
air on the valve seat even while closed. 
Referring to FIG. 1, in order to permit a wide speed range for the piston 
unit 12, a double acting drive mechanism 130 is provided to operate the 
exhaust valves 110. A pair of cam shafts 132, 134 are coupled by a chain 
drive 136 to the crankshaft 18. Identical double acting cam units 138 are 
mounted on the cam shafts 132, 134 at locations adjacent the exhaust 
valves 110. Each of the double acting cam units 138 includes a first cam 
140 with a first cam surface which is operative to open a respective 
exhaust valve 110 and a second cam, or preferably a pair of cams 142, 144 
with respective identical second and third cam surfaces which are 
operative to close the respective exhaust valve 110. The double acting cam 
units 138 are operative with identical cam follower units 146. 
Referring to FIGS. 1 and 6, each of the identical cam follower units 146 
includes connecting links 148, 150 which are pivotally connected at one 
end to pins 152. The pins 152 are rigidly mounted in and extend outwardly 
from an opposed pair of projecting lugs or ears 154, 156 that rigidly 
extend from opposite sides of the annular body 158 of the exhaust valves 
110. Each of the identical connecting links 148, 150 also include 
generally a oblong-shaped through hole 160 which has a minor diameter 162 
(FIG. 1) sized to mate with and receive the diameter of the cam shafts 
132, 134 passing through the hole 160. The hole 160 has a major diameter 
164 (FIG. 6) that is sized to permit the cam follower unit 146 to move 
through generally linear strokes or displacements corresponding to the 
motion of the exhaust valves 110 as they are opened and closed by the 
double acting cam unit 138. 
Referring now to FIGS. 3 and 6, each of the cam follower units 146 further 
includes shafts 166, 168 on which are mounted cam followers 170, 172, 174. 
Operation of the piston unit 12 rotates the crankshaft 18 which in turn 
causes cam shaft 132 to rotate thereby rotating the exhaust valve 
operating cams 140, 142, 144. As the cam shaft 132 rotates the cams 140, 
142, 144, respective cam surfaces 176, 178, 180 maintain continuous 
contact with the cam followers 170, 172, 174, respectively. However, 
because of the elliptical shape of the cams 140, 142, 144, the cam surface 
176 of cam 140 is operative to move the cam follower unit 146 to the left 
away from piston 22 and crankshaft 18 as viewed in FIG. 3, thereby opening 
the exhaust valves 110. During a further rotation of the cam shaft 132, 
the relationship between the cam surfaces 178, 180 of the eccentric cams 
142, 144, respectively are effective to move the cam follower unit 146 to 
the right toward piston 22 and crankshaft 18 as viewed in FIG. 3, thereby 
closing the exhaust valves 110. The positive double action of the 
operation of the exhaust valves 110 permits the piston unit 12 to operate 
over a wide speed range without any change in the magnitude of the forces 
required to operate the exhaust valves 110 as is the case with the use of 
compression springs to close the exhaust valves. Consequently, the torque 
required to actuate the exhaust valves is minimal which facilitates 
starting of the compression ignition piston unit 12 with the gas turbine 
unit 36. 
While the invention has been set forth by a description of the preferred 
embodiment in considerable detail, it is not intended to restrict or in 
any way limit the claims to such detail. Additional advantages and 
modifications will readily appear to those who are skilled in the art. For 
example, the operation of cam 140 and cams 142, 144 to respectively open 
and close the exhaust valves 110 may be reversed so that the cam 140 and 
cams 142, 144 respectively close and open the exhaust valves 110. In 
addition, the pair of operatively duplicate cams 142, 144 may be replaced 
by a single cam. Further, the pivotal connection of the connecting links 
148, 150 onto the trunion pins 152 may alternatively be a rigid 
connection. Even though eight cylinders are shown in FIG., 1, any number 
of pairs of opposed cylinders may be used. While the double acting cam 
unit is described as being preferable with the compound engine disclosed 
herein, the double acting cam unit can be used to operate the exhaust 
valve of any two-stroke uniflow scavenge engine unit. 
The invention therefore in its broadest aspects is not limited to the 
specific details shown and described. Accordingly, departures may be made 
from such details without departing from the spirit and scope of the 
invention.