Abstract:
The piston rod of the tandem-piston engine was modified to include a crosshead that slides in a crosshead guide machined into the lower part of the charger piston for better lateral support. The lower surface of the motor piston was changed to a simple conical surface allowing increased vertical thickness of the central region with a thinner outer region. A ceramic piston crown was clamped in compression by a piston bezel to the motor-piston base at a ledge with a conical upper surface having its vertex at the center of its base so thermal expansion will effect a sliding contact. A piston ring was used at the upper edge of the motor piston so the uppermost cylinder wall will be polished to prevent heat absorption. An intake port was given a flame arresting diffuser; and its entrance was blocked by the piston valve to prevent ignition in the valve cylinder.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
   Provisional Application No. 60/701,066 filed Jul. 21, 2005 by Ronald Dean Noland titled Tandem-Piston Engine 

   STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT 
   Not applicable 
   BACKGROUND OF THE INVENTION 
   My invention relates to internal-combustion engines of the type employing a motor piston that reciprocates in a cylinder while turning a crankshaft for the extraction of motive power from the expansion of hot gases produced by the combustion of a compressed fuel-air charge that is supplied under pressure to a combustion chamber by a charger piston. A type fitting into the broad classification of two-stroke pump-compression cycle internal-combustion engines. 
   Currently, most heat engines in use for mechanical power generation are of the four-stroke cycle type that was patented by Alphonse Beau de Rochas in 1862 and adapted for manufacture and practical use in 1876 by Nicolaus August Otto. These engines have been very successful for over a century and they are still the standard of comparison. Their mode of operation appears quite logical; but they do not make maximum use of the heat that is generated by the combustion of a fuel-air charge. And the various two-stroke cycle gasoline engines in current use suffer from the same problem. In the four-stroke Otto cycle, a fuel-air mixture is drawn into a cylinder having a combustion chamber into which the mixture is compressed by a piston. The fuel-air mixture is then ignited, and the heat released from burning the fuel greatly increases the pressure of the resulting gases by raising them to a very high temperature. The extremely hot high-pressure gases then push the piston down to turn the crankshaft and do work. But paradoxically, in Otto cycle engines all parts of the combustion chamber must be kept reasonably cool. If they are not cooled, heat absorbed from the burning gases during the power strokes will increase the temperature of the parts until something becomes so hot that it will ignite the fuel-air charge during the induction stroke. Ignition then results in a backfire explosion through the intake manifold that prevents the successful completion of the cycle. But even before any part becomes that hot, there will be preignition during the compression stroke that is likely to cause power loss, detonation and overheating of the piston and valves that can destroy the engine. Since it is essential that the combustion chamber of these engines be kept cool, it is necessary to incorporate an air or liquid cooling system that usually adds substantially to their bulk and complexity and detracts from their reliability. However, the combustion and expansion phase of the cycle would be far more efficient if it were to be conducted using a very hot insulated chamber. Then the extremely hot high-pressure combustion gases would not as quickly transfer the heat, which produced and maintains the high pressure, to the internal parts of the combustion chamber. These include the cylinder head, the valves, the piston crown, the spark plug and the exposed part of the cylinder wall. 
   An answer to the problem might be to induct only air and then inject the fuel into it after it is compressed, as in a Diesel engine. This will solve the preignition problem; however, if the parts in the combustion chamber are allowed to become very hot, the air will be heated and expanded as it enters the cylinder, so less air will enter. This will reduce the mass of air available for combustion. Also, if the air is being heated by the chamber surfaces while being compressed, more work must be expended to compress it. The efficiency then is likely to get worse instead of better as was shown by the research in the Adiabatic Engine Program. 
   Since a cool chamber is needed for induction, but a hot chamber is needed for expansion, a new approach is necessary before the thermal efficiency can be raised to a much higher level. The dilemma can be resolved by making an engine that inducts and compresses its fresh charge in a separate cavity. Then, at the proper time, the fresh charge is forced into a combustion chamber, which is insulated and kept very hot, for combustion and expansion. This will slow the rapid loss of heat from the combustion gases so that the pressure does not drop as fast while the piston is being pushed down. If less heat is lost, a higher expansion ratio can then be used to extract more mechanical energy from the heat energy. 
   The only way that it is possible to increase the indicated thermal efficiency of internal-combustion piston engines is to reduce the amount of heat wasted. This is true regardless of whether the waste is from incomplete combustion, forced convection followed by conduction through the chamber surfaces or out the exhaust pipe following a low ratio expansion. Increasing the expansion ratio will be of limited benefit in four-stroke spark-ignition engines where purging of heat from the combustion chamber is practiced because it is intrinsically necessary. And it is unlikely that the mechanical efficiency will ever be much improved while a four-stroke engine requiring two revolutions per cycle, an extensive valve train and a mechanically pumped and fanned liquid cooling system is used. 
   Previously, an engine was devised and patented that has a tandem-piston arrangement in which the mode of operation is to induct a fresh fuel-air charge into a cavity that is opened between the lower surface of a motor piston and the upper surface of a charger piston as they descend in a cylinder. When the pistons rise, the fresh charge is compressed between the pistons in a space called the charger cavity for injection into the combustion chamber, which is located above the motor piston. Injection is through transfer channels running upward past the motor piston along the cylinder wall that are opened as the motor piston rises past them when it nears its zenith. Injection and combustion take place as the motor piston passes its top center position. The transfer channel openings are then closed by the lower motor-piston ring as it descends early in the power stroke while the upper surface of the charger piston is in tight full-surface contact with the lower surface of the motor piston. The patent number is U.S. Pat. No. 5,509,382; it issued in 1996, and it is titled:  Tandem - Differential - Piston Cursive Constant - Volume Internal - Combustion Engine . For simplicity, the engine that is described in U.S. Pat. No. 5,509,382 will now be referred to as a tandem-piston engine. 
   BRIEF SUMMARY OF THE INVENTION 
   Continuing design and engineering work on the tandem-piston engine has shown the need for several improvements, which have been made. 
   Investigation of heat balance measurements in four-stroke engines led to the derivation of a means for estimating the indicated thermal efficiency of internal-combustion engines. The indicated efficiency is found in terms of the heat losses as a function of the engine speed and the temperature of the interior combustion chamber surfaces by the following formula:
 
η=[1−(1 /E ) k−1   ]·e   −KALt  
 
   It shows that the standard formula, η=1−(1/E) k−1 , for estimating the adiabatic efficiency, η, of an engine; that is, when no heat is lost into the chamber surfaces, may be multiplied by an exponential cooling factor, e −KALt , to find the expected indicated efficiency of the engine when there is heat being lost into the interior chamber surfaces. 
   In the formula, E is the expansion ratio, k is the ratio of specific heats, and K is a cooling constant having a value of 3.8 for an engine having a cylinder with a bore of around 76 mm (2.99 in.) and a stroke of 75 mm (2.95 in.). A is a thermal coefficient. It is calculated by using the average temperature of the combustion gases during the power stroke, about 2000 kelvin, subtracting the average interior combustion chamber surface temperature, about 500 kelvin, and then dividing the difference by the average temperature of the combustion gases. This gives it a value of about 0.75 in a small four-stroke engine. But in an engine such as the latest version of the tandem-piston engine, which has hot insulated combustion chamber surfaces, its value will decrease to around 0.25. The square root of the engine speed in revolutions per second is L, and t is the time for the stroke in seconds. 
   Using the formula derived for estimating the indicated efficiency, it was found that the engine as shown in the earlier patent, which only had an 11.8 to 1 expansion ratio, would only have a brake efficiency of around 38% or less, partly because the expansion ratio was far too low. Also, the insulators separating the various components in the motor piston, the piston valve and the cylinder head were not thick enough to prevent the fast transfer of large quantities of heat away from the combustion gases. 
   Calculations showed that an expansion ratio in the neighborhood of 25 or 30 to 1 would be necessary for very high efficiency. A new version of the tandem-piston engine was designed in which the actual stroke of the charger piston was increased to 72 mm (2.835 in.). Then the stroke of the motor piston was increased to two-thirds of the actual stroke of the charger piston, 48 mm (1.89 in.), leaving 24.77 mm (0.975 in.) for the effective charger stroke because the motor piston leads the charger piston in phase by 6 degrees. An expansion ratio of 30 to 1 was obtained by adjusting the volume of the combustion chamber for high-pressure operation while also including the expansion volume of the piston valve. 
   Thermal calculations then showed that far more insulation would be necessary to prevent heat loss from the combustion gases into the cylinder crown, piston crown and valve crown. These were redesigned to incorporate slabs of medium density zirconium phosphate of 19 mm (0.75 in.) for the cylinder crown and 12 mm (0.472 in.) for both the piston crown and the valve crown. This material is a machinable ceramic that is rated for use up to 1811 kelvin in air. It has a very low thermal conductivity, 0.9 watt/meter-kelvin, and a very low coefficient of thermal expansion, 0.9 micro-meter/meter-kelvin, which gives it an ultra high thermal shock resistance. The low coefficient of thermal expansion also allows it to be held in compression by clamping during high temperature operation. The bezels that hold the piston crown and valve crown were made of titanium for mass reduction and to use its low thermal conductivity to reduce the loss of heat from the combustion chamber. 
   The ceramic piston crown has a ledge with a conical upper surface near its perimeter with its vertex at the center of the lower supporting surface. The bezel that holds the ceramic piston crown has a mating conical surface. Since the vertex of the conical upper surface emanates from a point at the center of the lower supporting surface of the crown, then thermal expansion by either the ceramic piston crown or the bezel will neither loosen nor tighten the hold along the conical surface. There will be sliding contact from expansion by either. The bezel is also firmly fitted and tightened against the ceramic crown so as to hold it in compression by its outer surfaces since the ceramic is much stronger in compression than in tension. It is believed that the thermal expansion in the ceramic crown, which will operate at very high temperatures, will exceed the thermal expansion of the bezel, which is expected to operate around 400 to 500 kelvin. However, since the expansion coefficient of the ceramic is very low, the bezel could perhaps expand more, in that event, the crown will still be held firmly at the conical ledge. The valve crown was similarly redesigned with a ledge with a conical upper surface having its vertex at the center of its lower surface although the angle of the cone where the valve bezel holds it is much steeper. 
   The motor piston was also modified to give it a simple conical lower surface instead of the compound lower surface that was flat with a conical rim, which was used previously. This allows it to be thick in the center but much thinner at the outer edge thereby providing it with sufficient strength but less mass while it also reduces the length required for the transfer channels along the cylinder wall. The single conical surface is also easier to match with a single conical surface on the charger piston to effect total charge transfer. 
   The upper motor-piston ring has parallel conical surfaces. This form is used at the top of the piston bezel to polish the cylinder bore along the entire exposed section because it is believed that less heat will be absorbed where it is specularly polished. It is intended that the piston bezel be made of titanium for less mass; however, if titanium should prove to be unsuitable, it may be made of steel. The top valve ring and the valve bezel are similar in design and purpose. 
   The piston rod was redesigned to include a crosshead on its lower end that slides in a crosshead guide that is machined into the lower part of a redesigned charger piston so as to provide direct lateral support adjacent to the wrist pin. A strong lateral force is developed at the wrist pin as the motor connecting rod converts the vertical force from the piston rod into the rotaiy motion of the crankshaft. The lateral force will now be transmitted straight through the crosshead, the crosshead guide, and the wall of the charger piston to the cylinder wall without generating a twisting moment through the central piston rod boss. The original cantilever piston rod design lacked direct support for the lateral force at the wrist pin and it was thought that the force and wear might be excessive between the piston rod and piston rod boss. The new design should also effectively remove most lateral force from the motor piston. This arrangement is an improvement over the use of a conventional crosshead guide, which would need to be mounted in the cylinder block or upper part of the crankcase below the extent of piston travel, and which if used, would entail a substantial increase in the height of the engine. It would also entail an increase in the length of the charger connecting rods that would make them long and spindly. Another advantage of using a crosshead guide that is mounted in the charger piston is that the travel of the crosshead along a crosshead guide that is mounted in the charger piston is reduced to only the difference in stroke between the motor piston and the charger piston. In the present design, the stroke of the motor piston is two-thirds of the stroke of the charger piston so the travel of the crosshead along the crosshead guide is reduced to only one-half as much as it would be if the crosshead were to be mounted in the cylinder block in the typical fashion while the sliding velocity of the crosshead along the crosshead guide is also reduced by half. Another advantage of having the crosshead guide in the charger piston is that it is being pulled downward by the charger crankpins from the charger connecting rods while the motor piston is pushing the motor crankpin down through the piston rod and the motor connecting rod. Therefore the force that is transferred to the cylinder wall from the reaction at the wristpin from the motor rod is being reduced by the reaction at the piston pins from pulling the charger piston down thereby reducing the frictional losses even further. When totaled, the frictional losses are thus reduced to a small fraction of those encountered when using a crosshead guide mounted to the cylinder block. 
   A comparison of the latest preferred embodiment of the tandem-piston engine with a four-stroke engine having the same size main cylinder shows the following differences in operation and calculated performance: 
   1. It inducts only about one-third as much fresh charge, but it does it every revolution instead of once for every two revolutions as in a four-stroke cycle engine. 
   2. It has insulated combustion chamber surfaces that operate at from 1000 to 1500 kelvin to prevent the absorption and wasting of as much heat from the hot combustion gases. 
   3. It completely burns its fuel unless there is excess fuel in the charge. 
   4. It needs only a small fan for external cooling instead of a heavy, bulky liquid cooling system. 
   5. It uses a 30 to 1 expansion ratio to transform nearly 60% of the heat released by combustion into indicated mechanical work. 
   6. It has a power stroke with every revolution so it needs only half as many cylinders for equally smooth power delivery. 
   7. Its power output is comparable for the same cylinder size. 
   8. It uses low octane gasoline. 
   9. Its maximum brake thermal efficiency was calculated at 50%. 

   
     BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS 
       FIG. 1  is a full section view from the front of the small tandem-piston utility engine with the crank angle at 180 degrees AMTC; crankshaft rotation is clockwise. 
       FIG. 2  is a full section view from the left side of the small tandem-piston utility engine with the crank angle at 180 degrees AMTC. 
       FIG. 3  is a half section top view of the engine after removing the cooling shroud, and a half section top view of the engine with the cylinder head removed. 
       FIG. 4  is a half section view of the engine from the top with the sectioning through the exhaust ports in the upper part of the cylinder block. It also shows a half section view of the engine from the top with the sectioning across the intake manifold, the main intake ports, the middle intake ports, and the valve intake ports. 
       FIG. 5  is a view from the rear of the rear crankcase cover of the engine. 
       FIG. 6  is a front view of the front crankcase cover of the engine. 
       FIG. 7  is a section view through the link rod viewed from the right side of the engine showing its connection to the motor connecting-rod shank and the piston valve. 
       FIG. 8  is a section view from the top of the engine through the wrist pin and its supporting crosshead and the crosshead guide that is machined into the lower part of the charger piston. 
       FIG. 9  shows the motor-piston assembly with a line indicating the point of origin at the center of the lower surface of the piston crown for the conical upper surface on the ledge near the perimeter of the crown that is used for clamping by the piston bezel. 
       FIG. 10  shows the valve-head assembly with a line indicating the point of origin at the center of the lower surface of the valve crown for the conical upper surface on the ledge near the perimeter of the crown that is used for clamping by the valve bezel. 
       FIG. 11  shows the engine at 180 degrees AMTC when exhaust blowdown from the closing cycle is ending while filling of the charger cavity with a fresh fuel-air charge is being completed through the main intake ports. 
       FIG. 12  shows the engine at 236 degrees AMTC when induction ends as the charger piston rises past the main intake ports; meanwhile, the rising motor piston and piston valve are forcing hot residual gases out of the exhaust ports. 
       FIG. 13  shows the engine at 313 degrees AMTC as hot residual gases are trapped when the exhaust ports are closed by the rising piston valve; meanwhile, the fresh fuel-air charge is being compressed in the charger cavity. 
       FIG. 14  shows the engine at 343 degrees AMTC when the lower motor-piston ring opens the nine transfer channel inlets so the compressed fuel-air charge can surge into the combustion chamber for ignition by the fiery residual gases. 
       FIG. 15  shows the engine at MTC when the charger piston continues forcing the fresh fuel-air charge through the transfer channels into the combustion chamber where it ignites upon entry into the blazing combustion gases. 
       FIG. 16  shows the engine at 17 degrees AMTC during the power stroke when charge transfer ends as the lower ring of the motor-piston closes the inlets to the transfer channels while the pistons are meeting in tight full surface contact. 
       FIG. 17  shows the engine at 68 degrees AMTC during the power stroke when the charger piston starts opening the upper intake ports so that a fresh fuel-air charge can be drawn into the charger cavity through the valve cylinder. 
       FIG. 18  shows the engine at 110 degrees AMTC when the piston valve opens the exhaust ports for exhaust gas blowdown while the power stroke ends; meanwhile, the fresh fuel-air charge enters the charger cavity via the middle intake ports. 
   

   DETAILED DESCRIPTION OF THE INVENTION 
   Referring now to the drawings in detail,  FIGS. 1 through 10 . The engine has a main cylinder  1  that is fitted with a motor piston  2  of the crosshead type having a ceramic piston crown  3  that is clamped to it by a piston bezel  4 . A conical upper surface on the piston-crown ledge  5  that has its vertex at the center of the lower surface of the piston crown is used by a mating conical lower surface on the piston bezel  4  to hold the piston crown  3  against a pistoncrown shim  10  that covers a titanium motor-piston base  6  and a locking-bolt  8 . A piston-rod shim  13  is used to adjust the spacing of the motor-piston base  6  on a hardened and tempered steel piston rod  14  while a base bolt  7  clamps them together, and the locking bolt  8  with its height adjusted by a locking-bolt shim  9  prevents it from loosening. The piston-rod shim  13  must be chosen of the correct thickness to assure that the upper surface of the charger piston  27  firmly contacts the lower surface of the motor-piston base  6  at 17 degrees AMTC when a lower motor-piston ring  12  closes the entrance to the transfer channels  115 . The motor piston  2  has an upper motor-piston ring  11  that has conical upper and lower surfaces. The lower motor-piston ring  12  has a conical lower surface. The piston rod  14  has a crosshead  15  at its lower end, and a wrist pin  16  passes through it. The wrist pin  16  is held in place by wrist-pin keepers  17 , a keeper bolt  18 , and a keeper nut  19 . 
   A motor connecting rod  20  has a motor-rod shank  21 , motor-crankpin insert bearings  22  and a motor-rod cap  23  that is held in place by motor-rod bolts  24 . The upper end of the motor-rod shank  21  has an inner wrist-pin bushing  25  for the center section of the wrist pin  16  to turn in while its outer ends turn in outer wrist-pin bushings  26  that are fitted to the lower end of the piston rod  14  so that the wrist pin  16  pivotally connects the piston rod  14  to the motor connectina rod  20 . 
   An aluminum charger piston  27  has a crosshead guide  28  machined into its lower part so that forces developed at the wrist pin  16  may be transferred through the crosshead  15  to the crosshead guide  28  in the lower part of the charger piston  27  and from it to the walls of the main cylinder  1 . The charger piston  27  has a central piston rod boss  31  that has a groove for a contracting compression ring  29  and another for a contracting oil-control ring  30  with central-boss oil drains  32 . On its outer surface, it is fitted with a first charger compression ring  33 , a second charger compression ring  34 , an upper charger oil-control ring  35  and a lower charger oil-control ring  36  with skirt oil drains  37 . Inner piston-pin retaining plugs  38  and outer piston-pin retaining plugs  39  are used to keep the piston pins  40  in place. The motor-piston base  6  has a simple conical lower surface and the charger piston  27  has a mating simple conical upper surface to give full contact with simplified machining while allowing the motor piston  2  to be thicker in the central part but thinner at its outer perimeter. 
   Charger connecting rods  41  have charger-rod shanks  42  with piston-pin bushings  43  pressed into their upper ends and charger-crankpin insert bearings  44  in their lower ends that are retained in place by charger-rod caps  45 , and these are held on by charger-rod bolts  46 . 
   A crankshaft  47  has a motor crankpin  48 , charger crankpins  49 , a front main journal  50 , a rear main thrust journal  51 , a front counterweight  52 , a rear counterweight  53 , a front shaft  54 , a rear-output shaft  55  and crankshaft oil ducts  56 , which are closed by oil-duct plugs  57 . A crankcase  58  is closed by a rear-output-shaft seal  59  that is held in a rear crankcase cover  61  using a rear cover gasket  60 . At the rear, the crankshaft  47  turns in an upper-rear-thrust-journal shell  62  and a lower-rear-thrust-journal shell  63  that are clamped to the crankcase  58  by a rear-thrust-journal cap  64  that is bolted on by journal-cap bolts  65 . At the front it turns in an upper-front-journal shell  66  and a lower-front-journal shell  67  that are held in place by a front-journal cap  68 . 
   Lubrication is facilitated by an oil-pump inner rotor  69  working inside an oil-pump outer rotor  70  that is fitted against an oil-pump back plate  71  that is held against a front crankcase cover  77  by back-plate bolts  72 . A front cover gasket  76  is used to seal the front crankcase cover  77  to the crankcase  58  while crankcase-cover bolts  75  hold it on and a front-shaft seal  78  closes it. Oil-passage plugs  73  facilitate the machining of oil passages and an oil-pipe o-ring  74  seals the outlet. 
   A flywheel  79  is positioned by a flywheel key  80  and held securely by a flanged flywheel nut  81 . 
   A piston valve  82  bas an aluminum valve body  83  with a valve shank  84  and a valve head  85  on which are fitted a second valve ring  87 , a third valve ring  88 , and a fourth valve ring  89 . A top valve ring  86  is fitted to a valve bezel  97 . The valve body  83  is fitted with a valve-oil-ring upper half  90  and a valve-oil-ring lower half  91  and it has valve oil drains  92  below them. The valve body  83  has a body flange  93  at its upper end, and valve-head shims  94  are used to mount the valve head  85  on it at the proper height where it is held in place by valve-head screws  95 . A valve crown  96  is clamped down with its lower surface against a valve-crown shim  98  by a valve bezel  97  that bears against a conical upper surface on a valve-crown ledge  99  with its vertex at the center of the lower surface of the valve crown  96 . The piston valve  82  reciprocates in a valve cylinder  100 . 
   A link rod  101  is fitted with a gudgeon-pin bushing  104  and is attached by a gudgeon pin  102  to the valve body  83  where it is secured by gudgeon-pin retaining plugs  103 . At the other end of the link rod  101 , it is fitted with an inner-link-pin bushing  106  and attached to the motor-rod shank  21 , which is fitted with outer-link-pin bushings  107 , by a link pin  105  that is secured by link-pin keepers  108 . 
   A cylinder block  109  sits on cylinder-block shims  110  and is held to the crankcase  58  by cylinder-block bolts  111 . At its upper end there is a combustion chamber  112  and an exhaust passage  113 , and it is closed at its upper end by a cylinder crown  114 . Also there are 9 transfer channels  115  at its upper end, one of which has an ignition recess  116  where a sparkplug  117  is located. An ignition coil  118  provides a spark for ignition to start the engine. After the engine is running, recompressed hot gases provide ignition of the fresh charge as it enters the combustion chamber. 
   An intake manifold  119 , which has a manifold cover  120  and a carburetor mounting flange  121  having carburetor mounting studs  122 , surrounds the cylinder block  109 . Valve intake ports  123  allow part of the fresh fuel-air charge to be drawn into the valve cylinder  100  from which it initially travels through upper intake ports  125  having flame arresting diffusers into the charger cavity  124 . The upper intake ports  125  are blocked at their entrance in the valve cylinder  100  by the piston valve  82  during the period that their exit in the main cylinder  1  is exposed to hot combustion gases when the motor piston  2  is in the lower part of its stroke in order to prevent the possibility of ignition of the fresh fuel-air mixture in the valve cylinder  100 . As the charger piston  27  proceeds downward, the fresh charge travels through middle intake ports  126  into the charger cavity  124 . When the charger piston  27  is in the lower part of its stroke, the main intake ports  127  allow the final part of the fresh fuel-air charge to be drawn into the charger cavity  124  directly from the intake manifold  119 . After ignition and expansion of the fuel-air charge in the combustion chamber  112 , hot combustion gases leave through exhaust ports  128  flanked by exhaust-pipe studs  129 . 
   A cylinder head  130  with a cylinder head liner  131  houses the cylinder crown  114  while head bolts  133  hold them tight against a head gasket  132  and the cylinder block  109 . A cooling shroud  134  for directing air circulated by the flywheel  79  is held to the cylinder head  130  by shroud bolts  135 . 
   An oil pan  136  is the reservoir for lubricating oil that is drawn in through the oil intake  143  of the strainer bottom  142  that is held to a strainer cover  141  by a strainer bolt  146  having a spacer  145  with a strainer nut  147 . The oil then passes through a strainer screen  144  and an oil pipe  137 , which is clamped to the strainer cover  141  by a clamping sleeve  140 , having an oil-pipe flange  138  that is held to the crankcase  58  by a flange bolt  139 . The oil pan  136  has an oil drain plug  148 , and an oil pan gasket  149  provides the seal while it is clamped to the crankcase  58  using oil-pan bolts  150 . 
   In  FIGS. 11 through 18 , which show the engine with the pistons in each of eight key positions, engine rotation is given in degrees AMTC (after motor top center; that is, the number of degrees the crankshaft  47  has turned in a clockwise rotation after the motor piston  2  was at its top center position). Illustration of the cycle starts at 180 degrees AMTC. A fresh fuel-air charge is inducted into the charger cavity  124  that opens between the motor piston  2  and the charger piston  27  as they descend. As the pistons move upward, the charge is compressed and at −17 degrees AMTC (+343 degrees AMTC) the transfer channels  115  leading to the combustion chamber are opened as the lower motor-piston ring  12  rises past their lower edges. The compressed fuel-air charge then surges into the combustion chamber  112  where ignition takes place as it contacts the hot recompressed residual gases from the previous cycle that are trapped inside and are much hotter than the ignition temperature of the fuel. Injection continues until +17 degrees AMTC when the lower motor-piston ring  12  closes the entrances to the nine transfer channels  115  as the motor-piston  2  moves downward past their lower edges while the conical upper surface of the charger piston  27  meets the conical lower surface of the motor-piston base  6  in firm total contact. An extremely high ratio expansion then transforms about 55% to 60% of the heat energy put into the gases by combustion into indicated mechanical work. And after subtracting friction and pumping losses, around 45% to 50% of the heat input is expected to be available in the brake mechanical work output. 
   In the preferred embodiment design shown, the engine has an 80 mm (3.15 in.) bore, and the actual stroke of the charger piston  27  is 72 mm (2.835 in.), but the motor piston  2  has a stroke of 48 mm (1.89 in.) and it leads the charger piston  27  in phase by 6 degrees so the strokes must be subtracted geometrically. The difference in phase gives the charger piston  27  a net effective stroke of 24.77 mm (0.975 in.) compared to an actual stroke of 48 mm (1.89 in.) for the motor piston  2 . However, the volume of the charger cavity  124  is reduced by one-ninth from a piston rod  14  that runs through it for mounting the motor piston  2 . And the expansion volume is increased by the displacement of a piston valve  82 . Thus, the volume of the expansion chamber totals to about 2.64 times the volume of the charger cavity. 
   The engine is lubricated by motor oil that is carried in its oil pan  136  as in most four-stroke engines, but it has a power stroke during each revolution so only half as many cylinders will be required for the same smoothness of power delivery. Elimination of the timing gears, camshaft, poppet valves, water pump, cooling fan and the extra revolution per cycle helps in keeping friction and pumping losses per revolution at about the level as four-stroke engines of equal power even though the expansion ratio is far greater. The friction and pumping losses were calculated at 35.0 joules per revolution at 3000 RPM for the tandem-piston engine. This is about the same as the losses per revolution for a four-stroke engine of the same power. 
   Because of the extremely hot conditions in the combustion chamber  112 , it is assumed that there will be no appreciable amount of incomplete combustion unless there is excess fuel in the charge. 
   The volume of the combustion chamber  112  at contact, that is when the conical upper surface of the charger piston  27  is tightly against the conical lower surface of the motor-piston base  6  with injection of the charge complete, is 13 cm 3  ( 0 . 793  cu. in) so as to bring the peak pressure to around 7350 kPa (1066 psi), about the peak pressure in a conventional engine having a 12 to 1 compression ratio. And the total expansion volume when the motor piston  2  is at 186 degrees is 289 cm 3  (17.636 cu. in), which gives the engine a 22.2 to 1 expansion ratio from the contact position. However, the engine has a greater expansion ratio, 30 to 1, from its AMV (angle of minimum volume, which is at about 8 degrees after the motor piston  2  is at its top center position). It is expected that after injection begins at −17 degrees AMTC, the pressure will rise to a maximum level as the AMV is passed at about +8 degrees AMTC and continue for a few degrees until injection ceases when the conical upper surface of the charger piston  27  contacts the conical lower surface of the motor-piston base  6  at +17 degrees AMTC. Because the extraction of mechanical energy starts at the AMV, and the injection period ends about 9 degrees later, the expansion ratio will be calculated from the combustion chamber  112  volume at the AMV, but the adiabatic efficiency is treated as it is in a limited pressure cycle for a Diesel engine.
 
η=1−(1/ E ) k−1 [(αβ k −1)/((α−1)+ k α(β−1))],
 
   wherein α=the ratio of the maximum pressure to the compression pressure, and β=the ratio of the volume at the end of injection to the minimum chamber volume. The expansion ratio from the AMV of 30 to 1 would give the engine an adiabatic efficiency of about 1−( 1/30) 0.3 =64%. However since the engine operates in a limited pressure cycle, the figure is reduced to about 63%. 
   The exhaust starts opening at 110 degrees AMTC (70 degrees before the motor piston  2  is at bottom center). With this timing, the piston valve  82  then closes the exhaust ports  128  at 310 degrees AMTC and about a quarter of the residual gases that are left in the expansion volume after blowdown, 0.0726 liters (4.43 cu. In.), are trapped. The residual gases are then recompressed into the combustion chamber  112  so as to provide ignition of the next charge during its injection and to absorb part of the heat of combustion so as to buffer the temperature rise; thereby suppressing the loss of heat as well as the generation of nitrogen oxides. 
   The fresh charge at full load is considered to fill the charger cavity  124  to a pressure varying from 93 kPa (13.5 psi.) at 1000 RPM to 86 kPa (12.5 psi.) at 4000 RPM. A temperature of 293 kelvin (20 degrees Celsius) was used because the charge is cooled by fuel evaporation, and while the charger cavity  124  is fairly warm, it has no extremely hot valves, hot surfaces or residual gas. Assuming an intake manifold  119  temperature of 288 kelvin (15 degrees Celsius), this gives 90% volumetric efficiency at 1000 RPM or 83% volumetric efficiency at 4000 RPM. 
   In the tandem-piston engine, the fuel-air charge that is transferred from the charger cavity  124  into the combustion chamber  112  during about 35 degrees of crankshaft rotation, is only about one-third as large as the fuel-air charge that is inducted and compressed in a four-stroke engine of the same power. This is because a charge is transferred during each revolution, and the thermal efficiency is expected to be around 50% greater since less heat is lost by forced convection and the expansion ratio is far higher. 
   Even though the maximum pressure in this engine is expected to be around 7350 kPa (1066 psi) as in a 12 to 1 compression ratio conventional engine, it is not expected that it will have a requirement for high-octane gasoline. This is because any fuel-air mixture that has been transferred into the combustion chamber  112  is aflame as it enters, but any fuel-air mixture that is still in the charger cavity  124  is squeezed into an extremely thin layer between the conical upper surface of the aluminum charger piston  27  and the conical lower surface of the titanium motor-piston base  6  and neither surface is very hot. Preignition or detonation of the fuel-air mixture before it has left the charger cavity  124  should be effectively precluded by its intimate contact with the relatively cool metal surfaces; so it is expected that gasoline of very modest octane levels will suffice. 
   Because of the recapture and recompression in the combustion chamber  112  of a substantial volume of hot residual gas, which when recompressed will almost certainly be above the ignition temperature of the fresh charge, it is believed that throttling will generally not be necessary in the tandem-piston engine. It should be possible to control the load by varying the amount of EGR (exhaust gas recirculation) in the fresh charge, unless perhaps it is desired to operate it partially throttled when it is idling for smoother operation. This will allow most operation to be done with a stoichiometric mixture with EGR for cleaner operation. Maximum load is expected at about α=0.8 (F R =1.25) when more torque than economy is desired; however, it is expected that a stoichiometric fuel-air mixture at α=1 (F R =1) will give about 97% of the maximum load while burning only 80% as much fuel. 
   Since combustion of the fresh fuel-air charge is expected as it is injected, it is not necessary to keep the combustion chamber  112  cool as it is in the Otto cycle, so it may be operated at a very high temperature that is limited only by the insulating and refractory characteristics of the materials used for the cylinder crown  114 , piston crown  3  and valve crown  96 . These may be made of medium density zirconium phosphate, a machinable ceramic that is rated for use up to 1811 kelvin in air. This ceramic has a very low thermal conductivity, 0.9 watt/meter-kelvin, and a very low coefficient of thermal expansion, 0.9 micro-meter/meter-kelvin, which gives it an ultra high thermal shock resistance. The low coefficient of thermal expansion also allows it to be held in compression by clamping during high temperature operation. The piston bezel  4  and the valve bezel  97  that hold the piston crown  3  and valve crown  96  respectively are made of titanium for mass reduction and to use its low thermal conductivity to reduce the loss of heat from the combustion chamber. 
   Because the cylinder block  109  is externally cooled by air, its operating temperature is estimated to be around 500 kelvin. This will enable lubrication of the engine by petroleum oil in the usual manner. And with the piston crown  3  being machined from an insulating ceramic material, the upper motor-piston ring  11  and the lower motor-piston ring  12  will not be required to transmit nearly as much heat into the cylinder block  109  as the rings are required to do in conventional engines. This will allow smaller ring gaps and closer piston fitting while piston ring and cylinder wall life should be improved.