Abstract:
The fuel efficiency of an internal combustion reciprocating piston engine may be increased through selective secondary expansion of exhaust gas in the engine cylinders in order to recover exhaust gas energy which is otherwise wasted by cylinder blow-down at the end of the power stroke. Exhaust valve cam switching, intake valve deactivation, multiple exhaust valves, a specialized exhaust manifold arrangement and an exhaust gas diverter valve can be configured to enable a reciprocating engine to selectively operate in efficient eight stroke cycle compound mode when moderate engine power is demanded, then revert to conventional four stroke cycle non-compound mode operation when high engine power is demanded, without stopping the engine. For a road vehicle application, the benefit is substantially reduced highway cruising fuel consumption, while incurring minimal impact on engine weight, minimal impact on engine manufacturing cost, and no adverse impact on vehicle acceleration performance, hill climbing performance or trailer towing performance.

Description:
FEDERALLY SPONSORED RESEARCH 
     Not Applicable 
     SEQUENCE LISTING OR PROGRAM 
     Not Applicable 
     TECHNICAL FIELD OF THE INVENTION 
     The present invention improves upon the thermal efficiency of a four stroke cycle internal combustion reciprocating piston engine by means of selectively increasing engine volumetric expansion ratio. This increased engine expansion ratio recovers gas energy which is typically wasted during the engine exhaust stroke when the exhaust valve of a conventional engine opens and excess cylinder gas pressure equalizes with atmospheric pressure in a throttling, or blow-down process. The present invention affects the configuration of a reciprocating piston engine cylinder head, camshaft, combustion chamber valve timing and combustion exhaust gas manifold. The present invention targets any engine application where efficient operation over a variable range of engine power is required, especially those applications where an engine is required to operate at a moderate power output for a large portion of the engine operational duty cycle. The present invention is compatible with turbo-supercharged engines, but is especially beneficial when applied to normally aspirated and mechanically supercharged engines. Suitable applications include, but are not limited to, road and off-road vehicle propulsion engines, marine propulsion engines, auxiliary power unit engines, portable electric power generator engines and stationary electric power generator engines. 
     BACKGROUND OF THE INVENTION 
     An internal combustion engine configured to gain thermal efficiency by means of providing a expansion stroke longer than the compression stroke was invented by James Atkinson in 1882, and is known as the Atkinson cycle. In 2002, Toyota Motors employed the Atkinson cycle on their “Prius” gasoline-electric hybrid automobile by configuring the intake valve timing for late valve closure during the compression stroke. A disadvantage of the Atkinson cycle approach to enhance expansion ratio is that engine volumetric efficiency is reduced by the reduction in combustion chamber charge volume, which increases the weight per unit horsepower of the engine. For a vehicle application, the resultant increased engine weight leads to an increase in overall vehicle weight, which detracts from the goal of reducing overall vehicle fuel consumption. The present invention avoids this weight disadvantage by providing a means to selectively engage compound mode operation to increase expansion ratio, without affecting maximum engine power capacity when compound operating mode is de-selected. 
     The Curtiss Wright R-3350 turbo-compound radial airplane engine of the 1950&#39;s recovers exhaust gas energy, which would otherwise be wasted, by means of multiple power recovery turbines coupled through gearboxes to the engine output shaft. This solution minimizes the weight penalty of adding exhaust gas expansion cylinders, however the high cost of the power recovery turbines and their associated gearboxes has since precluded application of the power recovery turbine method for non-aviation use. The present invention avoids this turbine and gearbox cost penalty by configuring the engine&#39;s own cylinders to act as selective power recovery expansion cylinders. 
     The method of improving the efficiency of a piston engine through compounding by utilizing a second cylinder to further expand working gas exhausted from a first cylinder has been widely applied to piston steam engines since the early nineteenth century. This same well known multiple cylinder compounding principle is applicable to internal combustion engines. 
     U.S. Pat. Nos. 6,202,416 Gray, 5,199,262 Bell, 4,917,054 Schmitz, 4,250,850 Ruyer, 4,237,832 Hartig and 4,159,700 McCrum, describe multiple cylinder compounding applied to an internal combustion piston engine, similar to the principle traditionally employed for compounded steam engines, by dedicating some of the engine&#39;s cylinders as exhaust gas secondary expansion cylinders, and describe valve timing and cylinder motion timing methods to effect the transfer of exhaust gas from fuel burning fired cylinders to secondary expansion cylinders. The Bell patent also describes a separate crankshaft for the expansion cylinders, driven at twice crankshaft speed, for the purpose of reducing the required size of the expansion cylinders, reducing the weight penalty of the expansion cylinders as compared to the McCrum, Schmitz and Gray patents. However, even with the reduced expansion cylinder size, the addition of dedicated expansion cylinders according to these prior patents adds significant weight and bulk to the engine, which is counterproductive to the goal of reducing vehicle fuel consumption. The present invention differs from the Gray, Bell, Schmitz and McCrum patents in that, according to the present invention, the expansion cylinder can selectively change back and forth, while the engine is running, from functioning as an expansion cylinder to functioning as a conventional fired cylinder, whereas, according to the Gray, Bell, Schmitz and McCrum patents, the function of the expansion cylinders is fixed, such that they are unable to function as conventional fired cylinders. 
     In addition, the present invention provides a means to store a compressed charge of exhaust gas in an exhaust gas expansion chamber and an exhaust manifold until any such time as the expansion cylinder is ready to accept it, whereas, according to the Gray, Bell, Schmitz, Ruyer and McCrum patents, the stroke timing of the power and expansion cylinder pistons must be constrained to a specific relative crankshaft clocking angle in order to facilitate the transfer of the exhaust gas charge from the fired cylinder to the expansion cylinder. 
     The present invention requires at least two conventional poppet type exhaust valves in each expansion cylinder head, similar to the Hartig patent and similar to one variant of the Ruyer patent. However, in the case of these Hartig and Ruyer patents, one of the exhaust valves in each cylinder only functions when compound mode operation is selected and remains closed when all cylinders are firing. With an inactive valve occupying part of the cylinder head, there is less port area available for the functioning valves, which constrains port size, thereby detracting from volumetric efficiency and increasing engine specific weight. The present invention retains use of all the exhaust valves when all the cylinders are firing, thereby imposing no penalty on maximum power capacity. 
     U.S. Pat. Nos. 7,121,236 Scuderi and 6,789,514 Suh describe a split cycle engine configuration in which intake and compression takes place in a dedicated cylinder, then the compressed gas charge is transferred to a second fired cylinder in which the charge is burned, expanded and exhausted. Such a split cycle engine may be configured with a charge cylinder having smaller volumetric displacement than the combustion cylinder, thereby increasing expansion ratio and improving thermal efficiency. However, such a split cycle cylinder configuration incurs the same overall engine weight penalty as does the Atkinson cycle configuration because of the consequent reduction of total engine volumetric efficiency. The present invention differs from the Scuderi and Suh patents in that, according to the present invention, the expansion cylinder is not used for combustion while compound operating mode is selected, rather it is used for secondary expansion of completely burned combustion products gas provided by a separate fuel burning fired cylinder. 
     Cylinder deactivation is a known method of improving the efficiency of a spark ignition engine operating at moderate power output, as further described by U.S. Pat. No. 7,260,467, Megli, and SAE Technical Paper Jan. 26, 2003. General motors applied cylinder deactivation to production Cadillac car engines in 1981. With this known method, dis-engageable couplings of conventional design are provided as part of the valve train for some of the cylinders, which when selected, de-couple the affected valves from their respective valve drive cams, causing the affected valves to remain closed, thus preventing fresh charge air from entering or leaving the deactivated cylinders. Fuel to the deactivated cylinders is shut off by an automatic controller. In the traditional method, the deactivated cylinders repeatedly compress and expand a trapped air charge within the cylinder. The remaining engine cylinders function normally as fired cylinders. A consequent reduction in total air flow to the engine allows the intake throttle valve to be opened wider to maintain the same moderate amount of power output. The resulting reduction in charge air pressure drop across the throttle valve eliminates some of the charge air throttling losses, resulting in an estimated five to ten percent increase in part-power engine efficiency for this cylinder deactivation method, with no adverse affect on the engine&#39;s maximum power rating. Similar dis-engageable valve drive cam couplings comprise components of the present invention, and the present invention also gains efficiency benefits from reduced throttling losses, however, the present invention differs from the cylinder deactivation method in that the affected cylinder or cylinders do not function as deactivated cylinders, instead these cylinders actively expand combusted gas discharged from one or more fired cylinders. 
     U.S. Pat. No. 4,401,069, Foley, describes an improvement on the cylinder deactivation principle in which an axially moving camshaft can selectively shift between two cam profiles for each valve, without stopping the engine. Similar cam profile selectivity comprises a part of the present invention, however, like the Megli patent, the Foley patent facilitates only cylinder deactivation, whereas the present invention utilizes selective cam profile changing in order to facilitate the active expansion of combusted gas discharged from one or more power cylinders. 
     Individual working elements comprising the present invention may appear conventional, however, in the present invention these working elements combine according to a new operating principle which has not been contemplated in the prior art. 
     Notwithstanding the numerous prior systems contemplated for addressing efficiency losses associated with the conventional four stroke cycle engine, and in light of the increasing cost and scarcity of petroleum based motor fuel, there remains a need for a simple, low cost, and light weight method for recovering otherwise wasted exhaust gas energy during moderate engine power operation, without adversely affecting the engine&#39;s maximum power rating. 
     SUMMARY OF THE INVENTION 
     These and other needs are provided, according to the present invention, by an apparatus that is readily adaptable to any four stroke cycle internal combustion engine, whether a spark ignition engine, a compression ignition engine or a hybrid of the two, such as a homogeneous charge compression ignition engine. 
     Owing to the kinematics of the crankshaft and connecting rod mechanism serving to reciprocate the piston within the cylinder, the volumetric expansion ratio of a conventional normally aspirated engine is typically equal to the engine volumetric compression ratio. In the case of a compression ignition engine, the compression ratio is typically limited by the maximum peak combustion gas temperature and pressure that the cylinder can tolerate. In the case of a spark ignition engine, the compression ratio is typically limited by the need to avoid detonation of the fuel. Consequently, per conventional design practice, engine volumetric expansion ratio is a fixed parameter which cannot be altered without altering other functional aspects of the engine. The present invention implements a method to vary the volumetric expansion ratio of an engine while it is running, thereby optimizing either the fuel efficiency or the horsepower capacity, according to the amount of power being demanded from the engine. 
     An engine operating according to the present invention has two operating modes, compound mode selected and compound mode de-selected. When, as a result of high power demand on the engine, compound mode is de-selected, all cylinders function as power cylinders and the engine operates as a conventional four stroke cycle engine with normal volumetric efficiency and unrestricted power capacity, whereby exhaust gas discharges from each cylinder at the end of the power stroke directly to the exhaust collector conduit and overboard. When, as a result of moderate power demand on the engine, compound mode is selected, the function of one or more cylinders changes from that of power cylinder to that of expansion cylinder. This change is accomplished by changing the timing of the expansion cylinder valves and by diverting the flow path of exhaust gas after it discharges from the power cylinder. Exhaust gas displaced from the power cylinder passes through a pressurized exhaust manifold to the expansion cylinder where additional work is extracted from the exhaust gas before it is discharged to the exhaust gas collector conduit. When two expansion cylinder expansion strokes are completed for each power cylinder power stroke, engine expansion ratio is doubled and engine fuel efficiency is consequently increased. 
     When a mechanical supercharger is added to an engine to improve power capacity, normal practice is to reduce cylinder compression ratio to compensate for the increased temperature of the charge air produced by the supercharger so that charge detonation may be avoided at open throttle operation. An unavoidable effect of reducing compression ratio is a corresponding reduction of expansion ratio, which in turn increases combustion gas work losses at cylinder blow-down. Since the present invention effectively doubles cylinder expansion ratio, it is especially advantageous when applied to a supercharged engine. The present invention may also be applied to a turbo-supercharged engine, however, in this case, one preferred embodiment would utilize an exhaust gas waste gate valve to bypass the turbocharger turbine when compound mode is selected, so that exhaust gas may fully expand in the expansion cylinder without being subjected to back pressure from the turbocharger turbine. 
     Cylinder heads comprising four valves per cylinder and overhead camshafts are commonly used on currently manufactured engines. The present invention utilizes at least two exhaust valves at each expansion cylinder, which makes the present invention readily adaptable to currently manufactured engines. 
     When an engine operates according to the present invention with compound mode selected, exhaust gas flows through one expansion cylinder exhaust port at a time, instead of through both exhaust ports simultaneously, as is the case with compound mode de-selected. Although this results in a constriction of the total port area through which the exhaust gas must flow, there is no consequent deleterious effect on compound mode operation owing to this constriction. This is because compound mode is typically selected when only moderate power is required and the engine can operate at moderate speed. Consequently, the mass flow rate of exhaust gas moving through an individual cylinder port with compound mode selected at moderate speed is typically no greater than the mass flow rate of gas moving through each cylinder port when compound mode is de-selected, the engine is operating at high speed, and gas is flowing through both cylinder exhaust ports simultaneously. 
     Another benefit derived by the invention is reduced nitrogen oxide gas emission from the engine cylinders on engines equipped with direct fuel injection. By retarding the timing of the fuel injection event, peak combustion temperatures can be reduced, which ordinarily will adversely affect fuel consumption due to the loss of cylinder pressure over the power stroke, however, the present invention increases the overall expansion ratio, thus offsetting this loss of efficiency. In addition, the elapsed time duration of the overall expansion interval is extended, thus providing more time for the combustion process to be completed. 
     Still another benefit of the invention, as compared to conventional cylinder deactivation, is that cylinder and cylinder head temperatures are maintained at normal levels while compound mode is selected, whereas, when conventional cylinder de-activation is used, the cylinder and cylinder head tends to cool, which then can then lead to a period of increased cylinder gas emissions after the cylinder is re-activated, until the cylinder and cylinder head recover stable operating temperature. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a comprehensive isometric view of all relevant elements of a two cylinder embodiment of the invention. 
         FIG. 2  is a detail view from  FIG. 1 , comprising one element of the invention, showing functional elements of a first type of dis-engageable drive cam coupling, which, when selected, latches the coupling to the engaged position. 
         FIG. 3  is a detail view, similar to  FIG. 2 , showing functional elements of a second type of dis-engageable drive cam coupling, which, when selected, un-latches the coupling to the dis-engaged position. 
         FIG. 4  is a plan view of a two cylinder embodiment operating with compound mode de-selected. 
         FIG. 5  is a plan view of a two cylinder embodiment operating with compound mode selected. 
         FIG. 6A  comprises four side views of a two cylinder embodiment showing sequential motion of the pistons and valves, with compound mode selected, moving from zero through 270 degrees of crankshaft rotation. 
         FIG. 6B  comprises four additional side views of the same two cylinder embodiment as  FIG. 6A , showing sequential motion of the pistons and valves, with compound mode selected, moving from 360 through 630 degrees of crankshaft rotation. 
         FIG. 7  is a valve timing chart of a two cylinder embodiment operating with compound mode selected. 
         FIG. 8  is a plan view of a four cylinder embodiment operating with compound mode selected. 
         FIG. 9  is a valve timing chart of a four cylinder embodiment operating with compound mode selected. 
         FIG. 10  is a plan view of a four cylinder embodiment configured for two stages of compounding, operating with compound mode selected, and one cylinder functioning as an expansion cylinder. 
         FIG. 11  is a plan view of a one cylinder embodiment operating with compound mode selected. 
         FIG. 12  is a valve timing chart of a one cylinder embodiment operating with compound mode selected. 
         FIG. 13  is a comprehensive isometric view of all relevant elements of a two cylinder embodiment of the invention configured with powered actuators to move the cylinder port valves instead of rotable shaft driven cams and drive cam couplings. 
     
    
    
     LIST OF THE DRAWING REFERENCE NUMERALS 
     
         
           5  Power Cylinder 
           5   a  First Power Cylinder, Four Cylinder Engine 
           5   b  Second Power Cylinder, Four Cylinder Engine 
           5   c  Third Power Cylinder, Four Cylinder Engine 
           6  Power Cylinder Intake Valve Cam 
           7  Power Cylinder Exhaust Valve Cam 
           8  Fixed Cam Coupling 
           9  Intake Manifold 
           10  Throttle Valve 
           11  Crankshaft Connecting Rod Journal 
           12  Piston 
           13  Intake Port Fuel Injector 
           14  Expansion Cylinder 
           14   a  First Expansion Cylinder, Four Cylinder Engine 
           14   b  Second Expansion Cylinder, Four Cylinder Engine 
           15  Dis-engageable Cam Coupling 
           16  Lever Arm Type Cam Follower 
           17  Expansion Cylinder Intake Valve 
           18  Inner Plunger 
           19  Outer Cylinder 
           20  Compression Spring 
           21  Pin Segment 
           22  Hydraulic Piston 
           23  Solenoid Valve 
           24  Oil Distribution Manifold 
           25  Expansion Cylinder Intake Valve Cam 
           26  Expansion Cylinder First Exhaust Valve 
           27  Expansion Cylinder Second Exhaust Valve 
           28  Expansion Cylinder First Exhaust Cam 
           29  Expansion Cylinder Second Exhaust Cam 
           30  Expansion Cylinder Third Exhaust Cam 
           31  Expansion Cylinder Fourth Exhaust Cam 
           32  Exhaust Manifold 
           33  Power Cylinder Exhaust Port 
           34  Expansion Cylinder First Exhaust Port 
           35  Exhaust Gas Reservoir Chamber 
           36  Diverter Valve 
           37  Exhaust Collector Manifold 
           38  Cross Pin Spring 
           39  Intake Valve Camshaft 
           40  Exhaust Valve Camshaft 
           41  Lubricating Oil Port 
           42  Power Cylinder Intake Valve 
           43  Valve Actuator 
       
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     The present invention will now be described more fully hereinafter with references to the accompanying drawings, in which preferred embodiments of the invention are shown. This invention may, however, be embodied in many different forms and should not be construed as limited to the embodiments set fourth herein; rather, these embodiments are provided so that this disclosure will be thorough and complete and will fully convey the scope of the invention to those skilled in the art. Like numbers refer to like elements throughout. 
     Turning now to  FIG. 1 , power cylinder  5  has an intake valve cam  6  and exhaust valve cams  7 , and fixed couplings  8  between the cams and the valves. The engine has a conventional intake manifold  9  and intake throttle valve  10 . Per conventional practice, the crankshaft connecting rod journals  11  are clocked 180 degrees from each other such that when one piston  12  is at top dead center, the other piston is at bottom dead center.  FIG. 1  shows individual intake port fuel injectors  13 , but a conventional carburetor, a throttle body fuel injector, or fuel injection directly into the cylinder can also be used. 
       FIG. 1  shows that the engine has an intake valve camshaft  39  and an exhaust valve camshaft  40 , which rotate in a conventional manner through mechanical coupling to the crankshaft  11 . The expansion cylinder  14 , has a single dis-engageable coupling  15  for the intake valve  17 , functionally similar to that used on engines configured for conventional cylinder deactivation.  FIG. 1  depicts a lever type cam follower  16  situated between the intake valve stem  17  and the dis-engageable coupling  15 . However, there are many other known serviceable methods for implementing a dis-engageable coupling, depending on the configuration of a particular valve train mechanism, that are compatible with the present invention. For example, a dis-engageable coupling may be located on the cam follower lever itself, may be located between the valve stem and the cam or cam follower, or may be built into the tappet of a pushrod and rocker arm actuated overhead valve mechanism. 
       FIG. 2  is a cross section view of a dis-engageable coupling  15  taken from  FIG. 1 , illustrating functional elements of one type of conventional dis-engageable coupling which is compatible with the subject invention. Per conventional practice, the dis-engageable coupling can be combined with a hydraulic valve lash adjuster into a single cylindrical unit. As shown by  FIG. 2 , the coupling has an inner plunger  18  which can slide axially within an outer cylinder  19 . A compression spring  20  positioned between the inner plunger  18  and outer cylinder  19  extends the total height of the dis-engageable coupling assembly. A two piece cross pin, comprised of pin segment  21  and piston  22 , engages holes drilled crosswise through the inner plunger  18  and outer cylinder  19 . As shown on  FIG. 2 , when the coupling is selected, engine lubricating oil enters port  41 , exerts hydraulic pressure on piston  22 , and moves piston  22  and pin segment  21  to fully compress cross pin spring  38  so as to place the coupling in the selected, latched position, thereby preventing the dis-engageable coupling assembly from being compressed such that, as shown on  FIG. 1 , the second exhaust valve cam  31  can depress the second exhaust valve stem  27  by depressing the lever arm cam follower  16 . When the coupling is de-selected, there is an absence of lubricating oil hydraulic pressure at port  41 , consequently, cross pin spring  38  extends so as to move pin segment  21  and piston  22  to the de-selected, unlatched position. In this unlatched position, the ends of pin segment  21  align with the edges of the inner plunger  18  such that the coupling can absorb the motion of the second exhaust valve cam through compression of the inner plunger  18 , together with the pin segment  21 , against the compression spring  20 . The valve stem return spring is stronger than the compression spring  20 , resulting in the cam stroke being accommodated by compression of the coupling spring instead of by the valve stem spring, whereby the engine valve remains closed. 
       FIG. 3  shows an alternate configuration of a dis-engageable coupling, which utilizes similar components to that shown in  FIG. 2 , but instead of becoming latched when engine oil hydraulic pressure is applied to port  41 , the coupling becomes unlatched when engine oil hydraulic pressure is applied. 
     Accordingly, there are two types of hydraulic dis-engageable coupling applied to the subject invention. One type engages the latch as a result of hydraulic pressure, as shown by  FIG. 2 . The other type dis-engages the latch as a result of hydraulic pressure, as shown by  FIG. 3 . The present invention in this example utilizes both of these coupling types, depending on whether or not a particular cam is active or inactive during compound mode operation. As shown by  FIG. 1 , this permits a single hydraulic oil distribution manifold  24  to serve all of the dis-engageable couplings  15 , such that, when a single solenoid operated valve  23  opens to apply hydraulic pressure to the oil distribution manifold  24 , all coupling pin segment  21  and piston  22  latches move to place all four dis-engageable couplings  15  into the compound operating mode position; and when the solenoid operated valve  23  closes to shut off hydraulic pressure, the cross pin compression springs  20  then move all the cross pin segment  21  and piston  22  latches to place all four dis-engageable couplings  15  into the non-compound position. This example describes use of a hydraulic piston  22  to move a latching pin, however, an electro-mechanical device such as a solenoid or motor may be used in place of a hydraulic piston to achieve the same latching and un-latching effect on a dis-engageable coupling. 
     Referring now back to  FIG. 1 , when compound operating mode is selected, dis-engageable coupling  15  decouples an expansion cylinder intake valve  17  from an intake valve cam  25 , whereby the intake valve  17  remains closed. The expansion cylinder has a first exhaust valve  26  and a second exhaust valve  27 . The first exhaust valve  26  opens according to the position of either first exhaust cam  28  or second exhaust cam  29 . First exhaust cam  28  and second exhaust cam  29  each have a dis-engageable coupling  15  between the cams and the first exhaust valve  26 , to enable the opening sequence of the first exhaust valve  26  to be switched between that of first exhaust cam  28  and second exhaust cam  29 . When compound operating mode is selected, the first exhaust valve  26  couples with second exhaust valve cam  29 . When compound operating mode is de-selected, the first exhaust valve  26  couples with first exhaust cam  28 . The second exhaust valve  27  opens according to the position of third exhaust cam  30  and fourth exhaust cam  31 . Only fourth exhaust cam  31  requires a dis-engageable coupling linking it to the second exhaust valve  27 . When compound operating mode is selected, the second exhaust valve  27  couples with third exhaust cam  30  and fourth exhaust cam  31  simultaneously. When compound operating mode is de-selected, the second exhaust valve  27  couples only with third exhaust cam  30 . 
     As shown by  FIG. 1 , exhaust manifold  32  connects power cylinder exhaust ports  33  with the expansion cylinder first exhaust valve  26  port. Reservoir chamber  35  adds gas storage volume to the exhaust manifold  32 . This additional gas storage volume moderates the variation of gas pressure in the exhaust manifold  32  as the exhaust manifold  32  receives and expels individual charges of exhaust gas. If the engine is configured such that the exhaust manifold  32  by itself has sufficient volume, an exhaust gas reservoir chamber  35  is not required. When compound operating mode is selected, diverter valve  36  closes to prevent power cylinder  5  exhaust gas from escaping through the exhaust collector manifold  37 . When compound mode is de-selected, diverter valve  36  opens to allow free flow of exhaust gas from the exhaust manifold  32  to the exhaust collector manifold  37 . 
       FIG. 4  is a plan view showing gas flow through the two cylinder engine embodiment when compound operating mode is de-selected. Fuel and charge air are admitted to both cylinders,  5  and  14 . The exhaust diverter valve  36  remains open. Exhaust gas discharges freely from the ports of all four exhaust valves  33 ,  26  and  27 , into the exhaust manifold  32  and the exhaust collector manifold  37 , and from there to the atmosphere. Valve timing for the expansion cylinder  14  operates according to the conventional four stroke cycle, thus enabling both cylinders to produce maximum power according to the conventional four stroke cycle. 
       FIG. 5  is a plan view showing gas flow through the two cylinder engine embodiment when compound operating mode is selected. Charge air flows from the intake manifold  9  into the power cylinder  5  only. Fuel flow from the expansion cylinder injector  13  is shut off. The expansion cylinder intake valve  17  remains closed, preventing any fresh charge flow into the expansion cylinder  14 . Exhaust gas discharged from the power cylinder  5  flows through the exhaust manifold  32  and then into the expansion cylinder  14  through the port of the first exhaust valve  26 . Excess exhaust gas is stored temporarily in the reservoir  35 . Fully expanded exhaust gas discharges into the exhaust collector manifold  37  through the port of the second exhaust valve  27 . The diverter valve  36  is closed to prevent gas in the exhaust manifold  32  from escaping through the exhaust collector manifold  37 . 
       FIG. 6A  and  FIG. 6B  show the compound mode operating sequence of the two cylinder engine embodiment, with compound operating mode selected, in greater detail. For clarity, the continuously closed expansion cylinder intake valve  17  is not shown on  FIG. 6A  and  FIG. 6B , and only one of the two synchronized power cylinder exhaust valves  33  is shown. For further simplification,  FIG. 6A  and  FIG. 6B  disregard the effect of gas flow dynamics on valve timing, therefore, the optimum crankshaft position for opening and closing the valves will lead or lag those values presented by  FIG. 6A  and  FIG. 6B  to some degree, according to the desired operating speed of the engine and associated gas flow velocities. Referring now to  FIG. 6A , at zero degrees crankshaft angle, the power cylinder  5  has completed the power stroke, the power cylinder exhaust valves  33  begin to open and the expansion cylinder first exhaust valve  26  has opened. Between zero and 90 degrees, exhaust gas transfers from the power cylinder  5 , through the exhaust manifold  32  to the expansion cylinder  14  through a constant pressure displacement process. At 90 degrees, the first exhaust valve  26  closes. Between 90 and 180 degrees, the gas undergoes isentropic expansion in the expansion cylinder  14 , delivering a first increment of work to the crankshaft  11 . Meanwhile, the remainder of the exhaust gas transfers from the power cylinder  5  to the exhaust manifold  32  and to the reservoir  35 , temporarily raising the pressure of the combustion gas charge through an isentropic compression process. At 180 degrees, the power cylinder exhaust valves  33  close, trapping the exhaust gas charge in the exhaust gas manifold  32  and reservoir  35  under pressure. Meanwhile, the expansion cylinder second exhaust valve  27  opens, allowing the fully expanded gas charge to blow down into the exhaust collector manifold  37 . Between 180 and 360 degrees, the power cylinder  5  completes its intake stroke and the expansion cylinder  14  completes its first exhaust stroke. The expansion cylinder second exhaust valve  27  closes when the expansion cylinder piston  12  is a short distance below top dead center, compressing the remaining gas to a pressure approximately equal to the pressure of gas trapped in the exhaust manifold  32  and reservoir  35 . As shown on  FIG. 6B , at 360 degrees, the expansion cylinder first exhaust valve  26  opens. Since the pressure in the expansion cylinder  14  approximates the pressure in the exhaust manifold  32 , no gas energy is wasted by filling the expansion cylinder combustion chamber volume through blow-down of the exhaust manifold  32 . Between 360 and 450 degrees, the power cylinder  5  begins its compression stroke and exhaust gas transfers from the reservoir  35  and the exhaust manifold  32  to the expansion cylinder  14 , dropping in pressure as it does so. As a result, work which was done on the gas when it was compressed in the power cylinder  5  between 90 and 180 degrees is recovered in the expansion cylinder  14  between 360 and 450 degrees as this same gas charge undergoes initial isentropic expansion in the expansion cylinder  14 . At 450 degrees, gas pressure in the exhaust manifold  32  has returned to the same value it was between zero and 90 degrees, and the first exhaust valve  26  closes a second time. Between 450 and 540 degrees, the gas undergoes further isentropic expansion in the expansion cylinder  14 , delivering a second increment of work to the crankshaft  11 , and meanwhile the power cylinder  5  completes the compression stroke. At 540 degrees, the expansion cylinder second exhaust valve  27  opens, allowing the fully expanded gas charge to blow down into the exhaust collector manifold  37 . Between 630 and zero degrees, the expansion cylinder  14  completes its second exhaust stroke and the power cylinder  5  completes its power stroke. As before, the expansion cylinder second exhaust valve  27  closes early prior to the expansion cylinder piston  12  reaching top dead center so that expansion cylinder combustion chamber pressure approximates exhaust gas manifold  32  pressure when the expansion cylinder first exhaust valve  26  opens at zero degrees to repeat the cycle. 
     Each gas charge thus undergoes four conventional piston strokes in the power cylinder  5  plus four additional piston strokes in the expansion cylinder  14 , thereby comprising, in total, an eight stroke cycle over 720 degrees of crankshaft rotation. The first and second expansion cylinder work increments represent work recovered from the combustion gas that would otherwise be wasted by cylinder blow-down when an exhaust valve opens on an engine operating according to the conventional four stroke cycle. In this embodiment, the two isentropic expansion half strokes in the expansion cylinder  14 , from 90 degrees to 180 degrees and from 450 degrees to 540 degrees, together comprise a volume equal to the displacement of the power cylinder  5 , which effects a doubling of the engine expansion ratio as compared to an equivalent conventional four stroke cycle engine. 
       FIG. 7  shows the timing of the engine valves effected by the profile of the valve cams  6 ,  7 ,  25 ,  28 ,  29 ,  30  and  31  for the  FIG. 1  two cylinder engine embodiment.  FIG. 7  shows that at 180 degrees, there is no opening position overlap between the  FIG. 1  power cylinder intake valve cam  6  and power cylinder exhaust valve cams  7  when compound mode is selected. This is required during compound mode operation, because if both the power cylinder intake valve  42  and the power cylinder exhaust valves  33  were to be open at the same time, some of the compressed exhaust gas charge could be lost by reverse flow through the power cylinder exhaust valves  33 , through the power cylinder  5  and out the port for the power cylinder intake valve  42 . If valve overlap is desired to enhance power output for high engine speed, wide open throttle operation while compound mode is de-selected, a conventional camshaft phase shifting mechanism or a conventional variable cam lift mechanism may be added to modify the timing of the power cylinder intake valve  42 , or power cylinder exhaust valves  33 , such that power cylinder valve overlap can be de-selected when compound mode is selected. 
     Referring now back to  FIG. 5 , although combustion gas pressure in the exhaust manifold  32  varies throughout the engine cycle depending on the sequencing of power cylinder  5  emptying and expansion cylinder  14  filling, exhaust manifold  32  gas pressure during steady state operation with compound mode selected is never less than the pressure of the gas charge in the power cylinder  5  at the end of its power stroke. This aspect of the invention, in which the exhaust manifold  32  and reservoir  35  store a charge of gas from the power cylinder  5 , means that the present invention imposes no design constraints on an engine with respect to cylinder-to-cylinder piston stroke timing. As a result, the present invention may be applied to all manner of piston engine cylinder and crankcase configurations, including but not limited to, in-line, vee, opposed and radial configurations, to engines with any number of cylinders, and to engines with any cylinder firing order sequence. 
     The practical feasibility of the invention may be illustrated by calculating the additional work recovered by doubling the expansion ratio, as illustrated by  FIG. 1 , for a single operating condition with compound mode selected. In order to simplify the analysis, an air standard cycle with a constant specific heat ratio for air is assumed. 
     Assumptions:
     Air specific heat ratio, γ=1.4, constant   Air constant volume specific heat coefficient, C v =0.171 BTU/lbmF   Stoichiometric Air to Fuel ratio=14.7   Fuel Lower Heating Value=17,500 BTU/lbm   Sea level standard ambient pressure=14.67 psia   Sea level standard ambient temperature=59 deg F.   Intake throttling pressure drop=4 psig   Volumetric compression ratio, V 1 /V 2 =9:1   Spark ignition at top dead center (Otto cycle)   All fuel burns at top dead center after ignition   Friction work=w f =(0.1)(Indicated work with compound mode de-selected)
 
Operating States of the Cycle:
   

     0 Ambient 
     1 Start of compression stroke 
     2 Top dead center of compression stroke, prior to ignition 
     3 Top dead center of compression stroke, after ignition 
     4 Bottom dead center of power cylinder expansion stroke 
     5 End of expansion cylinder expansion stroke 
     6 Discharge of exhaust gas from the expansion cylinder
     Throttling of the intake charge:
 
 P   1   =P   0 −4=14.67−4=10.67 psia
 
T 1 =T 0 =59F=519R
 
Power Cylinder Compression:
 
 P   2   /P   1 =( V   1   /V   2 ) γ =9 1.4 =21.67
 
 P   2 =( P   2   /P   1 )( P   1 )=(21.67)(10.67)=231.3 psia
 
 T   2   =T 1( V   1   /V   2 ) γ−1 =519(9) (1.4−1) =1250 deg R
 
Ignition and Combustion of the Fuel:
 
 q =Fuel Lower Heating Value/Air-Fuel ratio=17,500 BTU/lbm/14.7=1190 BTU/lbm
 
 q=Cv ( T   3   −T   2 )=1190 BTU/lbm=0.171( T   3 −1250 R )
 
T 3 =8209 deg R
 
 P   3   /P   2   =T   3   /T   2 =8209/1250=6.57
 
 P   3 =( P   3   /P   2 ) P   2 =(6.57)231.3=1519 psia
 
Isentropic Expansion in the Power Cylinder:
 
 P   3   /P   4 =( V   4   /V   3 ) γ=( 9) 1.4 =21.67
 
 P   4   =P   3 /21.67=1519/21.67=70.1 psia
 
 T   3   /T   4 =( V   4   /V   3 ) γ−1 =(9) 0.4 =2.408
 
 T   4   =T   3 /2.408=8209/2.408=3408 deg R
 
Heat Rejected During the Power Cylinder Cycle:
 
 4   q   1   =Cv ( T   1   −T   4 )=0.171(519−3408)=−494 BTU/lbm
 
Net Indicated Work Produced During the Power Cylinder Cycle:
 
 w   p =1190−494=696 BTU/lbm
 
Isentropic Expansion in the Expansion Cylinders:
 
 P   4   /P   5 =( V   5   /V   4 ) γ =(2) 1.4 =2.639
 
 P   5   =P   4 /2.639=70.1/2.639=26.6 psia=11.9 psig
   

     Since P 5  is above ambient pressure, the exhaust gas was not over-expanded in the expansion cylinder, therefore work was done over the entire expansion interval, despite throttling of the intake charge.
 
 T   4   /T   5 =( V   5   /V   4 ) γ−1 =(2) 0.4 =1.320
 
 T   5   =T   4 /1.320=3408/1.320=2583 deg  R=T   6  
 
Heat Rejected from the Engine During Both the Power Cylinder and Expansion Cylinder Cycles:
 
 6   q   1   =Cv ( T   1   −T   6 )=0.171(519−2583)=−353 BTU/lbm
 
Total Net Indicated Work Produced During the Overall Engine Cycle:
 
 w   n   =q+   6   q   1 =1190−353=837 BTU/lbm
 
Work Recovered by the Expansion Cylinder:
 
 w   e   =w   n   −w   p =837−696=141 BTU/lbm
 
Work Delivered to the Crankshaft Output Coupling in Non-Compound Mode, Subtracting Friction Work:
 
 w   c   =w   p   −w   f =696−0.1(696)=626 BTU/lbm
 
     Additional work recovered by the invention in the expansion cylinder when compound mode is selected, as a percentage of the crankshaft output coupling work delivered if compound mode is de-selected:
 
Work recovered=( w   e   /w   c )(100%)=(141/626)(100%)=22.5%
 
     When efficiency gains from reduced throttling of an estimated five to ten percent are added to the calculated efficiency gains from increasing expansion ratio, the total engine fuel efficiency gain provided by the invention is: 22.5%+(5% to 10%)=27.5% to 32.5%. This calculated value of fuel efficiency gain is conservative owing to the simplified analysis. On an actual engine, the fuel does not burn instantaneously, instead it burns during a substantial portion of the power cylinder power stroke, resulting in a lower peak cylinder pressure, and consequently a higher exhaust gas charge pressure at the end of the power stroke. Therefore, the gas charge expanded in the expansion cylinder  14  will deliver more work energy than that calculated by this simplified analysis. Although determined by simplified analysis, this calculated engine fuel efficiency gain indicates the practical feasibility and usefulness of the invention. 
       FIG. 8  shows a four cylinder example of the present invention, illustrating how the present invention can accommodate any number of additional cylinders greater than the two cylinder example of  FIG. 1 .  FIG. 9  shows the timing of the engine valves effected by the profile of the valve cams for a four cylinder engine embodiment in which two cylinders function as power cylinders  5   a  and  5   b , and the other two cylinders function as selective expansion cylinders  14   a  and  14   b , thereby providing four expansion cylinder half-stroke work events per 720 degree engine cycle and yielding the same doubling of the engine expansion ratio as the two cylinder example of  FIG. 1 . 
     Although these example embodiments describe a doubling of the engine expansion ratio, the subject invention is by no means limited to increasing expansion ratio only by a factor of two. For example, a five cylinder engine can be configured with two power cylinders and three expansion cylinders, thereby yielding an expansion ratio of 5/2=2.5. When the same engine is re-configured with three power cylinders and two expansion cylinders, expansion ratio then becomes 5/3=1.67. Accordingly, the expansion ratio in compound mode may be configured as required to best suit the anticipated operating duty cycle of a specific engine configuration, depending on the overall number of cylinders comprising the engine. 
       FIG. 10  shows a four cylinder example of the present invention, illustrating how the present invention can accommodate multiple stages of selective compound operation in which one or more expansion cylinders are selected incrementally according to how much power is being demanded from the engine. Such progressive selection of the degree of compounding allows compounding to be useful over a wider range of engine power output than would be the case if only one stage of compounding is provided. Two stages of compounding may be obtained by configuring two exhaust gas diverter valves  36  and two exhaust gas reservoirs  35  in the exhaust manifold  32  as shown by  FIG. 10 . Engines with more than four cylinders can accommodate three or more exhaust gas diverter valves and three or more exhaust gas reservoirs, thereby further widening the useful power range that can be accommodated by selective compounding. 
       FIG. 11  shows how a single cylinder engine can be configured for selective compounding, by means of a single cylinder configured to alternate function between that of power cylinder and that of expansion cylinder, when compound operating mode is selected. Single cylinder compounding requires that the intake valve have two cams instead of one, and that the second exhaust valve  27  have two dis-engageable couplings instead of one. The camshaft rotates at one fourth crankshaft speed, according to an eight stroke cycle, over 1440 degrees of crankshaft revolution per cycle. Accordingly, each of the four exhaust valve cams has two lobes. When either sequential or direct fuel injection is used, the timing of the fuel injection events changes from one injection event every two crankshaft revolutions when compound mode is de-selected to one injection event every four crankshaft revolutions when compound mode is selected.  FIG. 12  is a valve timing diagram which shows that, during the exhaust stroke following the power stroke, all of the exhaust gas charge is compressed and stored in the exhaust gas reservoir chamber  35 , subsequently the following two expansion strokes deliver one increment of work each to the crankshaft  11 . Because initial pressure at the beginning of the first expansion stroke is higher than it is for the beginning of the second expansion stroke, the first exhaust valve  26  is timed to open prior to the piston reaching midstroke position for the first expansion stroke and opens after the piston reaches midstroke for the second expansion stroke. This produces an expansion ratio for the first stroke greater than two and an expansion ratio for the second stroke less than two, which makes the cylinder pressure at the end of both expansion strokes equal, thus minimizing blow-down losses at the end of both expansion strokes. The total swept volume for the two expansion cylinder half-strokes equals the total swept volume of the cylinder when it is acting as a power cylinder, which effectively doubles the engine expansion ratio as compared to operation in four stroke cycle mode when compound operating mode is de-selected. The cylinder configuration shown on  FIG. 11  may be applied to individual cylinders of a multiple cylinder engine in order to effect multiple stages of selective compounding, thereby widening the engine power output range over which compounding is useful, yielding a benefit similar to that provided by the alternate configuration shown by  FIG. 10 . 
       FIG. 13  shows that the intake and exhaust valves,  17 ,  42 ,  33 ,  26  and  27 , may be directly actuated by hydraulic or electromechanical actuators  43  instead of by conventional camshafts, cams, cam follower levers and dis-engageable couplings, while providing all of the variable valve timing characteristics shown by  FIG. 1 . The advantage of hydraulic or electromechanical valve actuation is reduction of mechanical complexity by elimination of the camshafts, camshaft drive mechanisms, cam followers and dis-engageable couplings. This advantage trades off against the disadvantage of the cost and weight of electronic or hydraulic power supplies and associated controls for the valve actuators  43 . 
     Many modifications and other embodiments of the subject invention will come to mind to one skilled in the art to which this invention pertains, having the benefit of the teachings presented in the foregoing descriptions and the associated drawings. Therefore, it is to be understood that the invention is not to be limited to the specific embodiments disclosed, and that modifications and other embodiments are intended to be included within the scope of the appended claims.