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منقول Brief Introduction to the InsuLated PuLse Engine

This personal research project describes a thermally efficient concept for combusting fuel in an internal combustion engine. It explores adiabatic engines, a usually dormant science that was last active for several years after the 1979 oil crisis. This concept is not like published adiabatic engines which expel superheated combustion gasses into an exhaust duct for the post-processing of energy. Instead, this "cold adiabatic engine” concept allows combusted gasses to adiabatically expand and cool before exiting the combustion chamber.

This is a work in progress. This draft is a conceptual paper, not a technical paper, as it contains intuitive approximations and primitive constructions which require refinement. The images exist as props to showcase the concept, not to suggest a best design practice. I’m not an engine designer and I don't claim this concept works. I only seek to learn why this concept may or may not work, perhaps find the flaw or fallacy which is understood by industry, and publish the results in this paper. This engine concept can be fabricated using century-old technology, and similar concepts have certainly been studied and dismissed, but the findings are not readily available. Technical critique and information on similar experiments is welcome.

Brief Introduction to the Insulated Pulse Engine

The "insulated pulse-combustion engine", abbreviated "insulated pulse engine" or "IPC engine", is a low volumetric efficiency internal combustion engine concept which combusts fuel at high thermal efficiency. The IPC engine sequence is thermodynamically quite simple, applying comparatively pure forms of: 1) isentropic (adiabatic) compression, 2) isochoric (pulsed) heat addition, 3) isentropic (adiabatic) expansion, and 4) isobaric heat rejection.

The P-V diagram of an ideal Otto cycle (not shown, but readily searchable) indicates the Otto (gasoline) engine also has a purely isochoric heat addition process, however it is well recognized the Otto engine more accurately incorporates a two-stage heat addition process which starts out isochoric and transitions to isobaric, with the isobaric segment significantly increasing volumetric efficiency while significantly decreasing thermal efficiency. The IPC engine concept principally differs from the Otto engine in that it:
1. Optimizes isentropic compression (much the way a Diesel engine does),
2. Eliminates the isobaric component of heat addition (much the way HCCI engine prototypes do),
3. Significantly extends isentropic expansion (much the way an Atkinson engine does), and in doing so,
4. Enables practical thermal insulation of the combustion chamber (much the way adiabatic engine prototypes did, except the IPC engine is able to use cheap durable insulators).


The combustion chamber of the IPC engine isselectively insulated using two economical Fe60Ni40 alloy inserts to minimize heat energy loss to a cooling system. Combustion initiates and is consumed rapidly near top dead center (TDC), permitting adiabatic cooling of combustion chamber gasses through the entire expansion cycle. The expansion cycle is extended beyond convention to extract additional heat energy from the combusted gasses, further reducing average combustion chamber temperatures to minimize stress on the thermal insulators, eliminating the need for a cooling system and resulting in an exhaust that is comparatively cool and pressureless.


Conventional emissions control devices don't work with low temperature exhaust gasses, so the IPC engine stratifies fuel to locally combust in a region of the combustion chamber specifically shaped to support efficient, clean combustion. Fuel stratification additionally permits throttling the homogenous fuel-air equivalence ratio within the highly reactive range of 0.40-0.80 to assure a rapid, complete combustion reaction while retaining a useful power band. A fuel-air equivalence ratio below 1.00 represents the deviation of a stoichiometric ratio toward fuel-lean.

Due to stratification constraints, the crankshaft RPM limit is defined by the combustion reaction rate of the selected fuel. Operating above the RPM limit of the selected fuel promotes incomplete combustion. When operated at or below the RPM limit, the IPC engine combusts cleanly with minimal need for emissions controls. For investigative purposes, gasoline is assumed the fuel used in the IPC engine, but diesel, propane, ethanol, ammonia, or most any conventional motor fuel is applicable to the IPC engine.


This website introduced a hypothetical 3.2 liter in-line 4-cylinder 4-stroke IPC engine with twin counter-rotating crankshafts on 04Sep2009 that generated an expected 50 HP at 4000 RPM. This website partly introduced a hypothetical 3.2 liter in-line 4-cylinder 2-stroke IPC engine on 08Apr2010 that generatedsimilar power. Much was learned about engine structures as design boundaries were stretched in these early models. Current activity is to reintroduce these engines in a more compact and practical "single crankshaft" form. For investigative purposes, the cylinder bore diameter of the IPC engine in this paper is 100mm and the piston stroke is 100mm.

When compared with naturally-aspirated Otto and Diesel engines at full throttle, a similarly displaced IPC engine at full throttle consumes roughly an eighth of the fuel each combustion event. This is based on the observation that HCCI prototype engines at full throttle consume a fourth of the fuel that Otto or Diesel engines of similar displacement consume at full throttle, and only half the piston stroke of the IPC engine is used during the compression cycle. The IPC engine is expected to have almost twice the peak fuel efficiency of Otto and Diesel engines, and will therefore generate roughly a fourth of the horsepower of similarly displaced Otto and Diesel engines at full throttle and similar RPM. Due to the increased thermal dwell of each combustion event, the IPC engine at reduced RPM is expected to have greater than twice the fuel efficiency of Otto and Diesel engines at reduced RPM.

The cylinder displacement requirements of an IPC engine are roughly four times that of Otto and Diesel engines at equivalent horsepower and RPM, but the cost, weight, and space requirements of the IPC engine assembly remain comparable due to a reduction in need for cooling, muffling, and emissions control components. Since mechanical friction is a variable which correlates more closely to generated horsepower than to displacement, and since the IPC engine is constructed using methods which emphasize reduction of mechanical friction, windage friction, and port pumping losses, friction generated within the IPC engine is comparable to friction generated within equivalently powered Otto and Diesel engines.
Fig10 – Introductory image of a deluxe version of the 3.2 liter, 50 horsepower, 2-stroke IPC engine concept, the deluxe version containing intake ports in the cylinder bore and exhaust poppet valves in the head. Twin crankshafts, as shown, reduce piston skirt friction and piston noise, but are found to undesirably increase engine cost, size, and weight. A single crankshaft is now the preferred direction for IPC engine modeling. The next IPC engine to be presented here will be an intermediate version of the 2-stroke which, because it requires no poppet valves or camshafts, is expected to be a simple, lightweight, low friction, low maintenance, and very low emissions engine. Both a tuned exhaust (shown in brown) and an untuned exhaust have been modeled for investigative purposes. The next IPC engine models will employ the untuned, free-flow exhaust, since a tuned exhaust adds cost but provides negligible performance benefit in an engine with a cool, pressureless exhaust.



Volumetric Efficiency and Thermal Efficiency

There is an ongoing effort to improve fuel mileage in motor vehicles. In the last half century, fuel mileage improvements from internal combustion engines have most often resulted from increased volumetric efficiency (i.e.: increased peak horsepower per unit volume of cylinder displacement), not increased thermal efficiency. Fuel mileage gains have come from a deliberate matching of small displacement engine to large vehicle such that an engine issimply tasked to operate within a more thermally efficient segment of its operating range.


Higher volumetric efficiency in modern engines does not indicate improved thermal efficiency in an engine. For example, an older 80 horsepower 2 liter engine and a modern 160 horsepower 2 liter engine will likely provide about the same fuel mileage in a particular small car application.

Small displacement engines with high volumetric efficiency operate at higher combustion chamber temperature and pressure and higher RPM than do similarly tasked large displacement engines, reducing combustion chamber surface area and reducing exposure time in which each combustion event can lose heat energy to a cooling system. These conditions keep a 160 horsepower 2 liter engine within a more thermally efficient segment of its operating range when matched to a large vehicle, leading to better fuel mileage than achievable with a 160 horsepower 4 liter engine in the same large vehicle.

Cooling System Efficiency Losses

Internal combustion engines incorporate a cooling system to quickly remove heat energy absorbed by combustion chamber metals after each combustion event. This removal is necessary, since chamber metals would otherwise attain the average temperature of the combustion chamber gasses, a temperature too hot in Otto and Diesel engines for sustainable engine operation. Heat energy conducted through the combustion chamber metal into the cooling system represents a significant reduction in the thermal efficiency of an engine, particularly at low RPM when the dwell time for each combustion event is longest.

Following the oil crisis of 1979, internal combustion engine manufacturers around the world began developing “adiabatic engine” prototypes which contained thermally insulated ceramic combustion chambers in an attempt to improve engine thermal efficiency without sacrificing volumetric efficiency. Thermally insulating the combustion chamber reduced, and sometimes eliminated, the need for a cooling system, thus retaining a larger fraction of combustion heat energy for mechanical work output. In order to retain volumetric efficiency, these adiabatic engines were designed to combust with a conventional low heat release rate. This low heat release rate superheated the combustion chamber gasses before expelling them into the exhaust duct for the purpose of energy recovery through turbocompounding and other post-processing methods.

Experimental results on three published ceramic adiabatic engine projects can be reviewed in SAE technical papers 810070 (1981), 820431 (1982), and 840428 (1984), with abstracts viewable at the SAE.org website and where the papers may be downloaded. Adiabatic engines of the 1980s operated under the most brutal conditions. Adiabatic engines provided improved fuel efficiency, but could not be made sufficiently reliable for commercial application.

The use of a ceramic material, or the use of any thermally insulating material, to insulate combustion chambers of internal combustion engines for the primary purpose of improving fuel mileage in vehicles has found minimal research interest in the industry since the conclusion of these experiments.

The ceramic adiabatic engine experiments being developed around the world in the early 1980s were much more secretive than the prolific adiabatic engine experiments of race car builder Smokey Yunick. Smokey's many different adiabatic engine prototypes, which became staples of popular automotive magazines from 1979 through 1984, similarly blended the stressful combination of high volumetric efficiency with increased thermal efficiency, and then added some racing magic. Material longevity may have been one reason his adiabatic engines failed to achieve commercial applicability, and streetability may have been another. The magazine articles didn't provide in-depth or follow-up analyses, but watchful engine manufacturers certainly got all the necessary detail.

Exhaust System Efficiency Losses

In both Otto and Diesel engines, and in the adiabatic engine experiments described above, combustion is engineered to progress gradually, beginning near TDC and continuing well into the expansion cycle. This low heat release rate allows a lot of fuel to gradually burn without exceeding the pressure limits of the combustion chamber, providing high volumetric efficiency and low thermal efficiency. Volumetric efficiency is high because the piston experiences high levels of combustion pressure through a significant portion of the expansion cycle. Thermal efficiency is low because the late burning fuel cannot adiabatically expand as many times as the early burning fuel. Late burning fuel causes large amounts of fuel energy to be lost to the exhaust in the form of heat and pressure.

Unfortunately, the quantity of fuel Otto and Diesel engines must consume to generate high levels of horsepower cannot be combusted entirely at TDC without exceeding the pressure limits of the combustion chamber, so Otto and Diesel engines intentionally reduce the burn rate to achieve high volumetric efficiency. As applied in the adiabatic engine experiments of the 1980s, this longer burn duration exposed ceramic combustion chamber surfaces to more heat energy, raising temperature gradients within the body of the ceramic. The lower heat release rate may have set up thermal gradient stresses within the ceramic which contributed to reduced ceramic durability. By contrast, HCCI engine prototypes in research laboratories today combust all fuel near TDC and none during the expansion cycle, and Atkinson engines extend the expansion cycle until useable combustion pressure is mechanically consumed. These latter two engines release less heat energy to the exhaust than do equivalently powered Otto, Diesel, and adiabatic engines.




The Thermal Efficiencies of Insulation, Combustion, Expansion, and Friction
Thermal efficiency in an internal combustion engine is comprised of the following four core efficiencies, the first three of which determine the "adiabatic efficiency" of an engine:
1. High "insulation efficiency" minimizes loss of combustion energy to a cooling system in the form of heat, and is driven by the thermal conductivity of the combustion chamber. If maximizing thermal efficiency is the primary goal, and if excessive heat is lost to a cooling system, the insulation efficiency must be improved. If improved insulation efficiency causes the combustion chamber material to overheat and fail, the average temperature of combustion chamber gasses through a full engine cycle must be reduced.
2. High "combustion efficiency" minimizes loss of combustion energy to the exhaust duct in the form of elevated exhaust temperature, and is driven by compression ratio, ignition timing, and combustion duration. If maximizing thermal efficiency is the primary goal, and if the temperature of combustion chamber gasses is excessive at the end of the expansion cycle, the combustion efficiency must be improved.

3. High "expansion efficiency" minimizes loss of combustion energy to the exhaust duct in the form of elevated exhaust pressure, and is driven by the expansion ratio. If maximizing thermal efficiency is the primary goal, and if the pressure of combustion chamber gasses is excessive at the end of the expansion cycle, the expansion efficiency must be improved.

4. High "friction reduction efficiency" minimizes loss of combustion energy to mechanical component friction and to air pumping friction within the engine.


Adiabatic efficiency of Otto and Diesel engines
Insulation efficiency is low in Otto and Diesel engines due to the high average temperature of combustion chamber gasses. The high average temperature of gasses necessitates the active cooling of combustion chamber materials to keep them at reliable operating temperatures, resulting in significant combustion energy loss to the cooling system. Combustion efficiency is low in Otto and Diesel engines because high volumetric efficiency does not allow full combustion at TDC without developing excessive cylinder pressure, creating the need for a combustion process with a gradual "low heat release rate", in which the latter combusting fuel expands at lower adiabatic efficiency than the earlier combusting fuel, the latter combusting fuel providing significant heat energy loss to the exhaust duct. Expansion efficiency is low in Otto and Diesel engines because the compression process and expansion process are conveniently of equal stroke length, resulting in significant pressure energy being released to the exhaust duct before it can perform work on the piston. It should be noted that the compression cycle and expansion cycle are independent functions and will seldom be of equal length in an engine optimized for high thermal efficiency.

Adiabatic efficiency of the ceramic adiabatic prototype engine
Insulation efficiency was high in the adiabatic engine experiments of the early 1980s, but combustion efficiency and expansion efficiency were both low. These “adiabatic engines” were, in effect, one-third adiabatic, not fully adiabatic. Thermal efficiency was high because the thermally insulating ceramic combustion chamber material prevented significant heat energy loss to a cooling system. Combustion efficiency was low because the fuel in these experiments burned with a low heat release rate in order to maintain conventional high volumetric efficiencies, with the latter burning fuel generating significant heat energy loss to the exhaust duct. Only the small portion of fuel burning near TDC combusted at high adiabatic efficiency. Expansion efficiency was low because significant pressure energy remained in the combustion chamber when the exhaust cycle began, with only a fraction of this energy recovered through post-processing. The combined result was a brutally hot expansion and exhaust process which provided some improvement in thermal efficiency over Otto and Diesel engines, but which excessively stressed the ceramic thermal insulators, rendering them unsuitable for commercial application.

Adiabatic efficiency of the HCCI prototype engine
Combustion efficiency is high in HCCI engines being researched around the world today, but expansion efficiency and insulation efficiency are both low. Combustion efficiency is high because the entire combustion reaction occurs at a “high heat release rate” near TDC, letting all fuel perform work on the piston from the start of the expansion cycle. Expansion efficiency is low because useable pressure remains in the combustion chamber when the exhaust cycle begins, resulting in the loss of significant pressure energy to the exhaust duct before it can perform work on the piston. Insulation efficiency is low in the HCCI engine, since an active cooling system is required to keep combustion chamber materials at a reliable operating temperature. The HCCI engine can also be considered a "one-third adiabatic" engine.

Adiabatic efficiency of the Atkinson engine
Expansion efficiency is high in Atkinson engines being produced today, but combustion efficiency and insulation efficiency are both low. Insulation efficiency is low in Atkinson engines, since an active cooling system is required to keep combustion chamber materials at reliable operating temperatures. Combustion efficiency is low, because fuel must combust at a thermally inefficient "low heat release rate" through a significant portion of the expansion cycle. Expansion efficiency is high, in that the expansion cycle is extended in stroke length beyond that of the compression cycle, allowing extraction of all useable pressure energy from the combustion chamber. Thisdefines the Atkinson engine as yet another type of “one-third adiabatic” engine.

Adiabatic efficiency of the Insulated Pulse conceptual engine
Insulation efficiency, combustion efficiency, and expansion efficiency are all high in the IPC engine, and the constructions described below will provide a notable increase in fuel efficiency over adiabatic, HCCI, and Atkinson engines while combusting cleanly, without need for pollution controls. The IPC engine is a true adiabatic engine construction, but to prevent confusion with established naming practice, the IPC engine is probably best called a “cold adiabatic engine”, since it transfers minimal heat to a cooling system and expels minimal heat energy and minimal pressure energy out the exhaust duct. It should be noted that, while the adiabatic efficiency is high in the IPC engine, "friction reduction efficiency" must be included to determine thermal efficiency.

Industrial Trends

The current emphasis of industry is to follow the path of high volumetric efficiency to improve fuel mileage in motor vehicles, however fuel mileage gains may become tougher to find assmall engines more routinely populate large vehicles. Atkinson engines, which achieve improved thermal efficiency through reduced volumetric efficiency, are found in some of today’s most fuel efficient cars. HCCI engine development programs, now popular in laboratories around the world, seek high thermal efficiency using a process which has low volumetric efficiency. These two enginessuggest the possibility that low volumetric efficiency may eventually lead the way toward significant improvements in engine thermal efficiency and fuel mileage.





Fig12 – Gas ported piston with low tension compression rings and reduced skirt contact. Presently, gas ported pistons are only applicable to racing engines, since any fuel entering the gas ports will remain uncombusted and be ejected as a pollutant. Since the stratified combustion chamber of the IPC engine does not allow fuel to enter the ports, gas ported pistons are not a pollution issue in an IPC engine, nor are the ports subject to clogging.


Exhaust Emissions

Exhaust emission concerns in the insulated pulse engine fall into four simplified categories:
1. Hydrocarbon (HC) exhaust emissions, representing fuel that is not combusted, are formed when fuel is in proximity of chilled combustion chamber crevicessuch as are found near the head gasket, upper piston ring, and intake valve seat.

2. Soot emissions, also known as particulate matter (PM) emissions, representing fuel that is 1/3 combusted, are formed when fuel is direct injected into the dense flame kernel of a compression ignition engine which has already consumed all adjacent oxygen.

3. Carbon monoxide (CO) emissions, representing fuel that is 2/3 combusted, are formed when fuel is combusted near chilled surfaces within the combustion chamber.

4. Oxides of nitrogen (NOx) emissions are generated when heat energy becomes unnecessarily high in the combustion chamber and the very stable 3-bond nitrogen molecule breaks apart.

The cause of exhaust pollution in internal combustion engines is complex but well understood, as are clean combustion methods which prevent pollution, and as are exhaust processing methods which remove pollution.

Constructions which promote clean combustion have been extensively adopted by the IPC engine, since the cool temperature of the IPC engine's exhaust renders many popular emissions control devices ineffective, as many depend on significant levels of exhaust heat to function. Combustion in the IPC engine issufficiently unique that some form of emissions control will likely be required, but emissions levelsshould be sufficiently low that incorporation of the needed controls will not significantly affect cost or thermal efficiency.

Exhaust emissions in the IPC engine concept can be characterized more clearly once the initial energy calculations for an IPC engine cycle are completed. Before this will happen, an acceptably mature IPC engine assembly will be modeled in CAD, followed by presentation of the energy equations and initial calculations, using the CAD assembly model to base the calculations. At present, there exists little formal math to support the claims and curiosities of this paper. This conceptual paper will become a technical paper once the initial energy calculations are posted to this website. The focus will then turn to formatting this website for presentation and continued concept characterization.

There may eventually be opportunity to run engine simulation software on the IPC engine concept which will move the energy, emissions, and stress calculations closer to simulating a physical engine, thereby providing an industry-accepted baseline which effectively proves or disproves the engine concept on paper. Whether proved or disproved, if this project gets to the point of running simulations, the simulation results will be published. If proven and published, expanded research interest might commence. If disproven and published, future investigation into this conceptual construction would be resolved in short order. To be sure, there is a lot to study, calculate, and learn from each stage of this development process.

Basic Description of the Insulated Pulse Engine

The insulated pulse engine is an ordinary reciprocating piston internal combustion engine which applies unthrottled air induction, direct fuel injection, spark ignition, and the following three unconventional functions, to achieve high thermal efficiency:
1. Rapid "pulse" combustion (like an HCCI engine).
2. Thermally insulated combustion chamber (like an adiabatic engine).
3. Extended expansion cycle (like an Atkinson engine).

These three unconventional functions combine to create an engine with both high thermal efficiency and low volumetric efficiency. Mechanical friction and windage friction take on greater significance in engines with reduced volumetric efficiency. The insulated pulse engine carefully manages friction reduction and cost to optimize thermal efficiency. Friction reducing methods to consider include:
1. Twin counter-rotating crankshafts eliminate piston side thrust friction.
2. With single crankshaft, a longer connecting rod reduces piston side thrust friction.
3. Anti-friction material applied to piston skirt reduces piston side thrust friction.

4. Reducing excessive piston skirt contact area reduces viscous friction.
5. Gas ported low-tension piston rings reduce piston sliding friction.
6. Minimize port flow volume and port flow resistance, avoid throttled induction.
7. Minimize crankcase windage using crankcase vacuum and strategic bulkhead vents.
8. Turbulence should mix fuel with air efficiently, not excessively.
9. Rolling contact bearings, where possible, consume less energy than friction bearings.

The resulting engine requires only an active oil cooler of ordinary capacity to support all cooling needs, does not require a muffler to function quietly, and exhaust gasses can be made sufficiently cool that the exhaust manifold may be molded of plastic.

Twin counter-rotating crankshafts have been investigated here as a potentially cost-effective method to reduce friction within the IPC engine. Twin crankshafts hold additional value in noise reduction. Though the IPC engine does not generate unusually high combustion chamber pressures, it does generate an unconventionally rapid rate of change in pressure near TDC, and this may promote an undesirable "slap" sound if side thrust is applied to the piston by the connecting rod due to application of a single crankshaft. The present modeling constructions suggest twin crankshafts may unacceptably add to the height, width, weight, and cost of the engine. Development has now shifted toward a more practical single crankshaft construction using a short steel connecting rod in combination with an aluminum block. Whether or not any of the friction reducing suggestions shown above are applied to the IPC engine is not actually material to the core engine concept. It is only necessary to observe that friction reduction will improve thermal efficiency, and cost-benefit studies will determine which suggestions merit incorporation.




Fig13 – The study of one form of a piston and rod assembly for a dual crankshaft construction indicates engine size and weight may be undesirably increased. Current direction targets a single crankshaft design.


1) Rapid “Pulse” Combustion

In the IPC engine, combustion initiates near TDC and is rapidly consumed near TDC, providing combustion with low volumetric efficiency and high thermal efficiency. The volumetric efficiency is low because a comparatively small amount of fuel will generate sufficient temperature and pressure near TDC to reach the limits which do not form NOx exhaust pollutants. Thermal efficiency is high because all of the combusted gasses adiabatically cool through the entire expansion cycle, greatly reducing the percentage of heat energy lost out the exhaust and lowering the average temperature of the combustion chamber. The ordinary methodsselected to achieve this high heat release rate are:
1. High compression ratio
2. Combustion chamber shaped to fully support efficient combustion
3. Fuel-air charge turbulently mixed prior to ignition
4. Combustion chamber turbulence present at time of ignition
5. Additional combustion chamber turbulence generated by combustion reaction
6. Fuel-lean equivalence ratio optimized for rapid, complete reaction
7. Spark ignition precisely controls combustion envelope
Complete combustion at TDC in the IPC engine does not generate destructive pressure, as there is an insufficient quantity of fuel in the combustion chamber during each combustion event to generate excessive pressure. Pressure and temperature limits in the IPC engine’s combustion chamber are not driven by structural limits, but are driven by the need to prevent the formation of NOx emissions during combustion. If temperature and pressure in the combustion chamber climb sufficiently high that the very stable 3-bond nitrogen molecule breaks apart and forms NOx emissions, then temperature and pressure must be readjusted below NOx-producing levels, since the IPC engine is intended to combust cleanly without pollution controls.


Engine misfire may occasionally cause an anomalousstoichiometric fuel-air mixture to combust at detonation pressures in the chamber. The IPC engine, like conventional engines, is constructed to handle this type of misfire condition.




Fig14 – Piston shown 9mm BTC, just as combustion chamber becomes fully insulating.

2) Thermally Insulated Combustion Chamber

The IPC engine thermally insulates the combustion chamber completely when the piston is within 9mm of TDC, and partly insulates when the piston is further than 9mm from TDC. Three reasons for insulating are to 1) increase thermal efficiency by minimizing heat energy loss to the cooling system during the hottest portion of the compression and expansion cycles, 2) to burn cleanly at TDC by assuring critical combustion chamber surfaces flash to higher temperatures during compression and combustion to prevent the formation of CO exhaust emissions, and 3) to bring the combustion chamber up to operating temperature as fast as possible to minimize HC and CO exhaust pollutants at cold engine start-up.

The preferred thermal insulating material in the IPC engine is an iron or steel alloy containing 40% nickel of 3mm thickness, with thermal conductivity of 10 W/m K at 200 degrees C. As a comparison, the thermal conductivity of cast A356-T6 aluminum is 130 W/m K at 200 degrees C with typical thermal gradient distance of 10mm between combustion chamber and cooling system, and compacted gray iron is 40 W/m K with typical gradient distance of 5mm.A ceramic popular in the adiabatic engine prototypes of the 1980s, with thermal conductivity of 2 W/m K, is reserved as a preferred insulator in a future state of development of the IPC engine concept. More research is needed before this ceramic, known as partially stabilized zirconia (PSZ), can be proven reliable. Powdered metal-ceramic composites and other combustion resistant thermally insulating materials may also find future value as a thermal insulator in the IPC engine.Discrete ceramic thermal insulators may also improve CO exhaust emissions, due to more significant heating of the combustion chamber surfaces during compression and combustion, but turbulence during compression and combustion is expected to heat thermally insulating nickel steel combustion chamber surfacessufficiently to minimize CO exhaust emissions. Selective application of commercially available ceramic film coatings to the nickel steel combustion chamber may alternately minimize CO emissions, if needed.

As indicated, there are thermally insulating materials which insulate better than the selected 40% nickel steel alloy, but ideal insulators will not substantially improve engine thermal efficiency. The selected nickel steel will perform nearly as well as an ideal thermal insulator at high engine RPM, and will only become significantly less efficient than ideal insulators at low engine RPM, when the heat energy of each combustion event has more time to be absorbed by combustion chamber material. Even at low RPM, the nickel steel combustion chamber remainssignificantly more thermally efficient than Otto and Diesel combustion chambers. The rate of fuel consumption is low at low RPM, and therefore application of non-ideal insulators has reduced significance.

3) Extended Expansion Cycle

The IPC engine incorporates an extended expansion cycle, much like an Atkinson engine, to let combustion energy perform additional motive work before being discharged to the exhaust. The extended expansion cycle further reduces average combustion chamber temperature, bringing the average combustion chamber temperature down to the level where a cooling system is not required at all, except perhaps when running at full throttle in hot ambient conditions. When cooling is required, excess heat is readily removed via an external oil cooling system of ordinary capacity.An expansion ratio value isselected which will assure expansion energy gains constructively exceed friction force losses through the entire expansion cycle. Fuel pricing may apply a market force which influences the final specification of the expansion ratio. Otto and Diesel engines have evolved such that the compression and expansion cycles are approximately matched in stroke length. The compression cycle and the expansion cycle are each driven by significantly different physical parameters and mathematical equations, and their lengths will seldom coincide if maximized thermal efficiency is a primary goal. In the 4-stroke IPC engine, the compression cycle is roughly half of a piston stroke and the the expansion cycle is roughly a full piston stroke. In the 2-stroke IPC engine the compression cycle is roughly one third of a piston stroke and the expansion ratio is roughly two-thirds of a piston stroke.

Fuel-Stratified Combustion Chamber

Two critical issues exist with the combustion process described for the IPC engine: First, complete full-throttle combustion at TDC with a stoichiometric mix of fuel and air generates destructive pressure levels. With gasoline selected as the fuel, calculations indicate that a fuel-lean equivalence ratio of no more than about 0.25 is required to prevent excessive cylinder pressure when all fuel combusts at TDC, but full-throttle equivalence ratios in this low range combust incompletely and generate significant CO exhaust pollutants, with part-throttle equivalence ratios becoming non-combustive, as demonstrated in HCCI engine prototypes. Second, with homogenously mixed combustion reactions, there exist chilled locations in the combustion chamber which don’t support efficient combustion, yet which contain fuel and air. Examples of these locations include the clearance volume between the O.D. of the piston and I.D. of the cylinder bore above the sealing rings, the surface of the head gasket exposed to the combustion chamber, and the junction adjacent to the intake valve and seat. HC pollution is created in these locations of a homogenously inducted combustion chamber.

A stratified combustion chamber can resolve both the destructive pressure and the incomplete combustion issues. By splitting the combustion chamber into two separate regions just prior to direct fuel injection, one region can be designed to contain only air, while the other generate a turbulent mix of fuel and air, permitting clean and fast combusting fuel-air equivalence ratios in the range of 0.40-0.80 while segregating "chilly" chamber featuresinto the air-only region of the combustion chamber. The fuel-air region can be optimally shaped to fully support combustion, and a transfer passage between the two regions can be designed to direct turbulent kinetic energy prior to ignition and then support efficient expansion of the combusting reaction.

The stratified combustion chamber for the IPC engine initially forms when the piston is within 12mm of TDC and becomes shaped for clean fast combustion only when the piston is within 0.5mm of TDC. Spark ignition is required to assure combustion occurs precisely within this positional constraint. The rate of the combustion reaction is driven, in part, by the selected fuel, the compression ratio, the fuel-air equivalence ratio, chamber turbulence, and engine RPM, and will require a specified length of time to burn completely and cleanly. This reaction time defines an engine RPM maximum which, if exceeded, will result in pollution emissions. The IPC engine operates with greatest thermal efficiency at or just below this RPM maximum, and it retains practical levels of thermal efficiency at lower RPM. A maximum RPM value of 4000 has arbitrarily been assigned to the IPC engine for investagive purposes.




Fig15 - Thermally insulating piston cap, made from investment cast 40% nickel steel.



Fig16 – Back view of piston cap and head dish showing insert-casting retention features.



Fig17 – Thermally insulating 40% nickel steel head dish with thermal conductivity of 10 W/m K. Poppet valves are omitted from some versions of the 2-stroke IPC engine, significantly simplifying this casting.


Detailed Description of the 4-stroke Insulated Pulse Engine

Insulated Combustion Chamber
The most significant feature of the IPC engine is the thermally insulated combustion chamber: The piston contains an insulating cap, and the cylinder head contains an insulating dish. This is the full extent of the insulation. The unique size and shape of the stratified combustion chamber has forced a reduction in the valve head diameter, requiring a multi-valve arrangement to retain low-restriction intake and exhaust flow. The shown configuration of eight valves per cylinder is primarily a research tool, intended to assure sufficient flow for early experimentation and characterization. The combustion chamber will later be optimized to reduce cost.

These two insulating components will be investment cast out of a nickel iron or nickel steel alloy chosen for low thermal conductivity, high temperature stability, and valve seat wear resistance. For valve seat wear resistance, there issome carbon added to permit localized induction hardening. One of these insulators is inserted into the die cast mold of an aluminum piston to keep reciprocating mass low, the other is inserted into the mold of a cast aluminum cylinder head to keep engine mass low. The head insulator combines the duties of valve seat, spark/injector mount, and head gasket sealingsurface.

The valves installed into this head assembly are made of an inexpensive stainlesssteel or nickel steel alloy chosen for comparatively low thermal conductivity and for tribological compatibility with the nickel steel valve seats. The cylinder bore and piston rings are cast of conventional engine materials to assure good lubricity and long life at minimal cost.

While a steel containing 40% nickel, processed to produce a microstructure with thermal conductivity of approximately 10.0 W/m K at 200 degrees C, is the preferred insulating material in the IPC engine prototype, it may not be the best in future applications. Discrete ceramic components may provide improved insulating performance, but are brittle and will require a significant development effort to reliably incorporate into this engine.

The preferred discrete ceramic used in the adiabatic engine experiments of the early 1980s is called partially stabilized zirconia (PSZ). SAE Technical Papers 820429 (1982) and 830318 (1983), with abstracts viewable at the SAE.org website and where the papers may be downloaded, discuss internal combustion engine uses for discrete PSZ ceramic components. The thermal conductivity of two preferred PSZ ceramic compositions is approximately 2.0 W/m K.

PSZ ceramic was not sufficiently durable in the adiabatic engine experiments to become commercially applicable, though it performed remarkably well considering the severity of testing. It is expected PSZ will perform quite reliably at the lower average combustion chamber temperatures and milder thermal gradients within the IPC engine, but it must be incorporated in a manner which applies minimal tensile loading, preferring compressive loading where loads must exist.

Discrete ceramic insulators will not likely improve thermal efficiency in the IPC engine by more than a few percent over the selected nickel steel, but the lower thermal conductivity of discrete ceramic insulators may improve thermal efficiency of the engine at lower RPMs, such that the increased dwell of each combustion event at lower RPMs does not result in significantly increased heat energy absorption into the combustion chamber material. Discrete ceramic insulators may also improve CO exhaust emissions, but it is expected that turbulence during compression and combustion will heat critical nickel steel combustion chamber surfaces sufficiently to prevent CO exhaust emissions.

A queued task is to determine whether the thermally insulating microstructure of the selected 40% nickel alloy is stable in the stressful operating environment of the IPC engine, or whether the selected alloy reverts to a more thermally conductive microstructure. Actual choice of insulating alloy is not a core issue, and will be determined when insulating material becomes the research focus for the IPC engine concept.

Because the cylinder of the IPC engine is made of conventional materials which are thermally conductive, the combustion chamber will only be fully insulating when the piston is within 9mm of TDC. With the brief combustion reaction near TDC, combustion chamber temperatures drop considerably by the time the cast iron cylinder bore issignificantly exposed to combustion chamber gasses, minimizing heat energy loss.

The combustion chamber predominantly insulates when the piston is within 9mm of TDC. The combustion chamber partially insulates when the piston is farther than 9mm from TDC. As the piston travels from TDC toward BDC, and while the piston remains closer than 9mm to TDC, the heat generated by the combustion reaction is almost entirely dedicated to applying force to the crankshaft, finding minimal opportunity to route heat energy to a cooling system.

The combustion chamber switches from predominantly insulating to partially insulating when the piston drops below 9mm from TDC, as a segment of thermally conductive cast iron cylinder bore starts to occupy a small portion of the combustion chamber'ssurface area. Combustion chamber gasses have adiabatically dropped in temperature by the time the thermally conductive cylinder bore surface becomes a significant percentage of the combustion chamber surface area, greatly reducing heat energy absorption into the cast iron cylinder.

The thermally insulating segments of the combustion chamber exist to reduce heat energy absorption, thereby preserving heat energy for mechanical work, and to assist with complete combustion to minimize pollutant emissions. The thermally conductive segments of the combustion chamber exist in order to circumvent the tribological development requirements associated with using thermally insulating materials as wear surfaces.




Oil Cooling
Since the thermally conductive cast iron cylinder bore cyclically forms a portion of the combustion chamber, it absorbs a small portion of the heat of combustion. The average cyclic temperature of the cast iron cylinder bore remains below that which requires active cooling. The thermally insulating portion of the combustion chamber slowly absorbssome of the heat of combustion and needs to transfer this heat away. The cooling method is managed by ordinary oil circulation within the engine. The oil circulation system assures all parts of the engine are lubricated as required, and all are kept at functional temperatures. Should the oil temperature climb to a designated upper limit, an external oil cooling circuit will activate. This remote cooling circuit includes a small radiator and blower fan. When wind and cold weather are present, the IPC engine is best suited to operate in an enclosure without ambient venting, to prevent overcooling.

Since the IPC engine can be operated in conditions where the oil temperature remains cool for extended periods (cold climates, short trips), the oil may become saturated with water and degrade. An oil heat exchanger can be incorporated adjacent to an exhaust duct, and exhaust gasses can temporarily be routed through the oil heat exchanger whenever oil is below a specified minimum operating temperature. Since reactive combustion energy does not contact the cylinder bore in an IPC engine, cylinder bore oiling requirements are not assevere as those in conventional engines in which a flame contacts the internal bore.










Stratified Combustion Chamber
The uniquely shaped combustion chamber of the IPC engine forms a small but significant volume between the O.D. of the piston and I.D. of the cylinder bore above the compression sealing rings. Thissmall cylindrical volume is not shaped to support efficient combustion, and will generate pollution emissions if fuel is allowed to occupy this volume. Similar inefficient volumes within the combustion chamber exist at the head gasket and valve seats.

Modern Otto cycle engines design the pistons to minimize the inefficient cylindrical volume above the compression sealing rings, and the few exhaust emissions forming in the existing small volumes are scrubbed clean by a catalytic converter. Minimizing this volume in an IPC engine requires a reduction of thermal insulation coverage, in order that the sealing rings can be located as close as possible to the compression end of the piston This can reduce thermal efficiency of the engine. Additionally, the IPC engine generates a comparatively cool exhaust when compared with Otto and Diesel engines, and conventional catalytic converters do not perform efficiently at these lower exhaust temperatures. For this reason, the IPC engine must take another approach to eliminating crevice-sourced pollutants.

The IPC engine is designed to prevent the creation of pollution in areas of the combustion chamber which don’t support efficient combustion, since it is designed to keep direct injected fuel out of these locations. The established way to keep fuel away from these locations is to operate as a Diesel cycle engine, spontaneously combusting direct injected fuel as it enters the combustion chamber, but Diesel engines intrinsically suffer from soot emissions, since fuel must be injected directly into the center of a dense flame kernel which has already consumed all adjacent oxygen. Diesel engines must remove soot pollution from the exhaust using a particulate burner, but a particulate burner does not function efficiently with the comparatively cool exhaust of the IPC engine.

The solution for preventing the creation of pollution emissions in the IPC engine is found in combustion chamber stratification. Combustion chamber stratification, in coordination with an insulated combustion chamber, pulse combustion, uniquely timed direct injection, and spark ignition, create a combustion environment which favors clean combustion and minimizes the creation of exhaust pollutants, minimizing the need for emissionscontrols.

The combustion chamber of the IPC engine isstratified only when the piston is located within 12mm of TDC. When the piston is farther than 12mm from TDC there exists only one region in the chamber. The stratified combustion chamber forms when the piston is at 12mm BTC, segregating into a perimeter squish region (called a "crevice chamber" in older images) which actively rejects fuel and a central combustion region (called a "central combustion chamber" in older images) which is optimized to mix injected fuel with air and combust cleanly. An annular transfer passage (called an "annular passage" or a "backfill passage" in older images) communicates between the two regions, transferring air toward the central combustion region as the piston rises above 12mm BTC, returning fully combusted gasses to the perimeter squish region as the piston falls to 12mm ATC. The annular transfer passage also provides a buffer at TDC which efficiently constrains the combustion reaction.

The perimeter squish region assists complete combustion: It keeps fuel away from combustion chamber features which do not efficiently support combustion. While the piston approaches TDC the perimeter squish region acts as an air pump which transfers air toward the central combustion region to turbulently mix injected fuel with air prior to ignition. Direct fuel injection begins when the piston is 8mm BTC and ends by 6mm BTC. The direct injector nozzles are aimed to inject fuel mass only into the piston pocket at the center of the central combustion region. The air pumping action actively constrains all direct injected fuel to the central combustion region, permitting selection of preferred fuel-air equivalence ratios in the range of 0.40 to 0.80 which combust most rapidly and cleanly, rather than the pollution-prone 0.15 to 0.25 equivalence ratio range which would occupy a homogenous IPC engine’s combustion chamber. Note that the volume of the perimeter squish region approaches zero at TDC, whereas the volume of the central combustion region approaches a finite value at TDC, creating an effective air pump directed from the perimeter squish region toward the central combustion region in the last 12mm before TDC.

The central combustion region isshaped to fully support combustion: The surface area of the central combustion region is comparatively low to assist a speedy combustion reaction. The insulated chamber surface heats up quickly during compression and combustion to assure fuel in close proximity to the insulated material combusts properly. The central combustion region isshaped to generate within itself a toroidal vortex as air is pumped in from the perimeter squish region, assuring all fuel is in motion to uniformly combust, the turbulence minimizing both cold and hot spots in the central combustion region which helps prevent pre-ignition.

The annular transfer passage acts to buffer combustion at TDC. As the combusting reaction heats up at TDC, it expands beyond the central combustion region. The combusting gasses efficiently spill into a segment of the annular transfer passage which fully supports combustion, while pure air residing within the annular transfer passage is pushed into the perimeter squish region. Only when the piston falls 0.5mm after TDC do combusted gassessignificantly occupy the annular transfer passage and begin to approach the perimeter squish region. By this time the combustion reaction has been consumed and concluded.

Any residual fuel that is not completely combusted when the piston falls to 0.5mm ATC will exit the combustion chamber as a pollutant. There is not a second opportunity to combust fuel that does not combust near TDC. If exhaust emissions are to be low, creviced features, such as valve seats and spark plug insulation recesses, are not permissible in the combustion area. The central combustion region at TDC is sized to be half the volume of the perimeter squish region plus the backfill passage at TDC, allowing a full throttle fuel-air equivalence ratio of 0.80.


Compression Ratio and Expansion Ratio
The IPC engine inducts unthrottled air, much like a Diesel engine. The IPC engine adiabatically pre-warms the induction charge during compression to just below the auto-ignition temperature of the fuel-air mixture, promoting rapid combustion when a spark is generated near TDC. This puts the dynamic compression ratio (DCR) at roughly 15:1. In flex-fuel configurations, the compression ratio is actively regulated to assure compression pressure remains just below the autoignition level as conditions change. This is accomplished by monitoring ignition reactivity and continuously servoing valve closure timing to suit.

The dynamic expansion ratio (DER) will be about 30:1 to minimize heat energy loss to the exhaust duct, much the way an Atkinson engine minimizes exhaust energy loss. The selection of 30:1 for the DER is based on the assumption that a peak combustion chamber pressure of 150 bar at TDC will not form oxides of nitrogen pollutants, and on the prevalence of predominantly diatomic gasses of the fuel-lean combusted charge obeying, to a first order approximation, the 150 bar / (30 ^ 1.4) = 1.3 bar equation. Mechanical friction drives a deviation from the 1.3 bar specification at BDC, though fuel prices may additionally influence the selected expansion ratio. An unconventionally large expansion ratio is chosen to extract virtually all useable heat and pressure from the combustion chamber before the exhaust valve opens, resulting in a comparatively cool and quiet exhaust cycle with minimal exhaust duct flow velocity.

The DCR can be referred to as the "compression ratio", and the DER referred to as the "expansion ratio". If the 4-stroke IPC engine has a 100mm piston stroke, the expansion cycle occupies 100mm of piston travel after TDC, and the 15:1 compression cycle begins 50mm BTC. The 15:1 compression ratio is independent of the 30:1 expansion ratio, and each can be adjusted as required.

The induction cycle for a deluxe version of the 4-stroke IPC engine occupies only the first 50mm of piston travel after TDC and the compression cycle occupies the final 50mm of piston travel before TDC. Combustion chamber pressure will drop as low as 0.50 ^ 1.4 = 0.38 bar in the period between the end of the induction cycle and the start of the compression cycle. The described induction cycle requires a valve train configuration with an unusually large camshaft base circle, in order to actuate the valves through such a small camshaft arc. An intermediate version of the 4-stroke IPC engine reduces cost and complexity by incorporating the Atkinson reversion cycle, in which the induction cycle occupies the entire 100mm of piston travel from TDC to BDC, and as the piston then rises from BDC the inducted air flows backward out the intake duct until the intake valve closes at 50mm BTC.



Fig25 – Perimeter squish region (formerly called the crevice chamber), central combustion region (formerly called the central combustion chamber), and annular transfer passage (formerly called either the annular passage or alternately the backfill passage), are formed at 12mm BTC.


4-Stroke IPC Engine Sequence

The deluxe 4-stroke IPC engine cycle includes twelve stages of operation including: 1) Intake, 2) Vacuum, 3) Rebound, 4) Compression, 5) Injection, 6) Turbulence, 7) Ignition, 8) Combustion, 9) Expansion, 10) Vacuum, 11) Rebound, and 12) Exhaust. The engine cycle includes the following sequence:
04mm ATC: Intake valve opens, drawing in unthrottled air, same as a Diesel engine.
50mm ATC: Induction cycle ends, intake valve closes.
51mm ATC: Cylinder begins pulling a vacuum as piston continues toward BDC.
100mm BDC: Combustion chamber drops to 0.50 ^ 1.4 = 0.38 bar pressure.
99mm BTC: Piston elastically rebounds off vacuum and is pulled toward TDC.
50mm BTC: Vacuum rebound ends, compression of inducted air begins.
49mm BTC: Inducted air begins adiabatically heating in combustion chamber.
12mm BTC: Combustion chamber transitions to become stratified.
09mm BTC: Combustion chamber becomes predominantly thermally insulating.
08mm BTC: Fuel is direct injected toward pocket at center of piston.
07mm BTC: Perimeter squish region pumps fresh air toward piston pocket, constraining fuel.
06mm BTC: Direct fuel injection ends.
05mm BTC: Air from perimeter squish region generates turbulence in central combustion region.
01mm BTC: Fuel and air homogenously mixed in central combustion region.
0.5mm BTC: Spark ignites fuel and air mixture, combustion progresses rapidly.
00mm TDC: Combustion reaction expands into annular transfer passage.
0.3mm ATC: Annular transfer passage forces pure air back into perimeter squish region.
0.5mm ATC: Combustion reaction extinguishes.
05mm ATC: Combusted gasses are adiabatically cooling in combustion chamber.
09mm ATC: Combustion chamber first exposes thermally conductive cylinder bore.
12mm ATC: Stratified combustion chamber transitions to become single chamber.
75mm ATC: Combustion chamber starts pulling a vacuum (low throttle only).
87mm ATC: Combustion chamber starts pulling a vacuum (mid throttle only).
99mm ATC: Combustion chamber pressure drops to 1.3 bar (full throttle only).
100mm BDC: Combustion chamber pressure or vacuum depends on throttle position.
99mm BTC: Expansion cycle ends, exhaust valve opens (full throttle only).
87mm BTC: Combustion chamber vacuum ends, exhaust valve opens (mid throttle only).
75mm BTC: Combustion chamber vacuum ends, exhaust valve opens (low throttle only).
04mm BTC: Exhaust valve closes.
04mm ATC: Intake valve opens, drawing in unthrottled air, same as a Diesel engine.

The deluxe 4-stroke IPC engine described above has a rather complex valve train. The deluxe 4-stroke IPC engine can be cost-reduced to employ a simpler, slightly less thermally efficient Atkinson reversion cycle. This intermediate version of the 4-stroke IPC engine follows the sequence:
04mm ATC: Intake valve opens, drawing in unthrottled air, same as Diesel engine.
100mm BDC: Induction cycle ends, Atkinson reversion cycle begins.
50mm BTC: Intake valve closes, Atkinson reversion ends, compression cycle begins.
49mm BTC: Fresh air begins adiabatically heating in combustion chamber.
12mm BTC: Combustion chamber transitions to become stratified.
09mm BTC: Combustion chamber becomes predominantly thermally insulating.
08mm BTC: Fuel is direct injected toward pocket at center of piston.
07mm BTC: Perimeter squish region pumps fresh air toward piston pocket, constraining fuel.
06mm BTC: Direct fuel injection ends.
05mm BTC: Air from perimeter squish region generates turbulence in central combustion region.
01mm BTC: Fuel and air homogenously mixed in central combustion region.
0.5mm BTC: Spark ignites fuel and air mixture, combustion progresses rapidly.
00mm TDC: Combustion reaction expands into annular transfer passage.
0.3mm ATC: Annular transfer passage forces pure air back into perimeter squish region.
0.5mm ATC: Combustion reaction extinguishes.
05mm ATC: Combusted gasses are adiabatically cooling in combustion chamber.
09mm ATC: Combustion chamber first exposes thermally conductive cylinder bore.
12mm ATC: Stratified combustion chamber transitions to become single chamber.
50mm ATC: Conventional expansion cycle ends, Atkinson expansion cycle begins.
100mm BDC: Atkinson expansion cycle ends, exhaust valve opens, exhaust cycle begins.
04mm BTC: Exhaust valve closes, exhaust cycle ends.
04mm ATC: Intake valve opens, drawing in unthrottled air, same as Diesel engine.



2-Stroke IPC Engine Sequence

A deluxe 2-stroke IPC engine incorporates an engine operating sequence summarized as follows:
1) Compression - 33mm BTC to 0.5mm BTC
2) Ignition – 0.5mm BTC
3) Combustion – 0.5mm BTC to 0.5mm ATC
4) Expansion – 0.5mm ATC to 67mm ATC
5) Induction - 67mm ATC to 90mm BTC
6) Exhaustion - 90mm BTC to 33mm BTC


A deluxe 2-stroke IPC engine includes intake ports on the cylinder bore and exhaust poppet valves in the head, and the operating sequence includes:
33mm BTC: Exhaust valve closes, compression of fresh air and some exhaust begins.
32mm BTC: Fresh air begins adiabatically heating.
12mm BTC: Combustion chamber transitions to become stratified.
09mm BTC: Combustion chamber becomes predominantly thermally insulating.
08mm BTC: Fuel is direct injected toward pocket at center of piston.
07mm BTC: Perimeter squish region pumps fresh air toward piston pocket, constraining fuel.
06mm BTC: Direct fuel injection ends.
05mm BTC: Air from perimeter squish region generates turbulence in central combustion region.
01mm BTC: Fuel and air homogenously mixed in central combustion region.
0.5mm BTC: Spark ignites fuel and air mixture, combustion progresses rapidly.
00mm TDC: Combustion reaction expands into annular transfer passage.
0.3mm ATC: Annular transfer passage forces pure air back into perimeter squish region.
0.5mm ATC: Combustion reaction extinguishes.
05mm ATC: Combusted gasses are adiabatically cooling in combustion chamber.
09mm ATC: Combustion chamber first exposes thermally conductive cylinder bore.
12mm ATC: Stratified combustion chamber transitions to become single chamber.
33mm ATC: Conventional expansion cycle ends, Atkinson expansion cycle begins.
66mm ATC: Combustion chamber pressure reaches 1BAR.
67mm ATC: Intake port becomes visible to combustion chamber.
68mm ATC: Vacuum forms and pulls fresh air into lower third of combustion chamber.
69mm ATC: Upper 67mm of chamber contains gasses with ¼ of oxygen consumed.
90mm ATC: Exhaust valves in head begin to open.
100mm BDC: Intake ports are fully visible to combustion chamber.
99mm BTC: Lower 33mm of combustion chamber contains air, upper 67mm contains exhaust.
90mm BTC: Intake ports in cylinder bore become blocked by rotating drum valve assy.
89mm BTC: Piston pushes combusted gasses in upper chamber into exhaust duct.
33mm BTC: Exhaust valves close, compression of fresh air and some exhaust begins.

An intermediate version of the 2-stroke IPC engine employs rotary drum valves for both induction and exhaustion, with intake and exhaust ports located on the cylinder bore, and the engine operates in a sequence summarized as follows:
1) Compression - 33mm BTC to 0.5mm BTC
2) Ignition – 0.5mm BTC
3) Combustion – 0.5mm BTC to 0.5mm ATC
4) Expansion – 0.5mm ATC to 67mm ATC
5) Simultaneous Induction and Exhaustion - 67mm ATC to 33mm BTC


This intermediate version of the 2-stroke IPC engine contains no poppet valves in the head. Instead, this intermediate 2-stroke IPC engine uses intake ports on one side of the cylinder bore, and exhaust ports on the opposite side of the cylinder bore. The intake ports are organized into an upper bank of ports and a lower bank of ports in which the rotary drum valve acts as a shutter and also acts as a blower. The exhaust side of the cylinder bore issimilarly configured, except the rotary drum valve assembly acts as a vacuum pump. Induction and exhaustion occur simultaneously, flowing across the combustion chamber with sufficient chaos that combusted gasses throughout the combustion chamber are substantially replaced with inducted air. The detailed operating sequence is as follows:
33mm BTC: Exhaust port sealed by piston ring, compression begins.
32mm BTC: Fresh air begins adiabatically heating.
12mm BTC: Combustion chamber transitions to become stratified.
09mm BTC: Combustion chamber becomes predominantly thermally insulating.
08mm BTC: Fuel is direct injected toward pocket at center of piston.
07mm BTC: Perimeter squish region pumps fresh air toward piston pocket, constraining fuel.
06mm BTC: Direct fuel injection ends.
05mm BTC: Air from perimeter squish region generates turbulence in central combustion region.
01mm BTC: Fuel and air homogenously mixed in central combustion region.
0.5mm BTC: Spark ignites fuel and air mixture, combustion progresses rapidly.
00mm TDC: Combustion reaction expands into annular transfer passage.
0.3mm ATC: Annular transfer passage forces pure air back into perimeter squish region.
0.5mm ATC: Combustion reaction extinguishes.
05mm ATC: Combusted gasses are adiabatically cooling in combustion chamber.
09mm ATC: Combustion chamber first exposes thermally conductive cylinder bore.
12mm ATC: Stratified combustion chamber transitions to become single chamber.
33mm ATC: Upper intake and exhaust ports first enter chamber but are shuttered closed.
50mm ATC: Upper intake and exhaust ports remain shuttered but are fully in chamber.
67mm ATC: Lower intake and exhaust ports first become exposed to chamber.
68mm ATC: Cross-flow of inducted and exhausted gasses begins in chamber.
69mm ATC: Upper ports begin to unshutter and begin to cross-flow.
90mm ATC: All intake and exhaust ports are now unshuttered and cross-flowing.
100mm BDC: Lower intake and exhaust ports fully visible to combustion chamber.
90mm BTC: Intake ports become blocked by rotary drum valve assembly.
89mm BDC: Piston pushes combustion chamber gasses out exhaust ports.
33mm BTC: Exhaust port sealed by piston rings, compression begins.

A simplified version of the 2-stroke IPC engine takes the previously described "opposing" intake and exhaust ports and places them on the same side of the cylinder block, with a single rotary drum valve assembly modified to provide both induction and exhaustion, and with the externally visible surfaces of the rotary drum valve assembly acting on the induction gasses and the internally visible surfaces of the rotary drum valve assembly acting on the exhaust gasses via a window in the drum. This results in a low-cost 2-stroke IPC engine.



Fig26 – Annular transfer passage (formerly called the backfill passage) has three times the volume of the perimeter squish region (formerly called the crevice chamber) at TDC.



Fig27 – 8 valves per cylinder is the simplest theoretical arrangement for the 4-stroke IPC engine. It should be remembered this is a theoretically ideal arrangement with application during early prototyping. Advanced computational analysis will cost-reduce this construction once the ideal arrangement is found to be valid. The intermediate and simplified versions of the 2-stroke IPC engine require no valves or camshafts, making for simple, low cost prototypes which may carry the economy of this combustion chamber directly into production.





Friction reduction
Mechanical friction and windage friction are significant issues in a low volumetric efficiency engine like the IPC engine. Presented here are some of the friction reducing methods being considered for this engine:

Twin gear-synchronized counter-rotating crankshafts can eliminate piston side thrust forces, eliminating skirt thrust friction and piston slap. Twin crankshafts are presently implemented in the IPC engine in order to learn about their characteristics, but the next stage of development will eliminate them due to cost, weight, and size constraints.



If twin counter-rotating crankshafts are chosen, resulting in the elimination of piston side thrust forces, the piston skirt can be sculpted to reduce contact surface area with the cylinder bore without concern for skirt wear, resulting in reduced piston skirt viscous oil friction losses.

In single crankshaft configurations, long connecting rods and anti-friction material applied to piston skirt will each reduce piston side thrust friction. When considering the short duration of high cylinder pressure after TDC, skirt friction need not be a significant issue in a single crankshaft version of the IPC engine.


Since the insulated pulse engine has high cylinder pressures for a very short percentage of crankshaft rotation, gas ported low-tension piston rings reduce piston sliding friction through the bulk of the piston travel while sealing tightly when pressure is high. The IPC engine’s combustion chamber is transitionally stratified and permits only pure filtered air to enter the piston’s gas ports, preventing localized HC exhaust emissions and also preventing gas port clogging. Presently, there are no production engines which can use gas ported pistons and achieve low exhaust emissions, limiting application only to racing engines. Since friction minimization is a primary consideration in the success of any low volumetric efficiency engine, the IPC engine may find gas ported pistons crucial to achieving commercial applicability.


As noted in the previous two paragraphs, the combination of short-duty skirt thrust forces and short-duty for high ring tension, may provide a unique opportunity to significantly reduce friction in the IPC engine.

Intake port friction reduction: Intake air in the IPC engine is unthrottled, as it is in Diesel engines. Throttled intake air, as found in Otto engines, consumes more pumping energy than unthrottled intake air.
Exhaust port friction reduction: There is minimal opportunity for “tuning” the exhaust with headers, since there isso little energy in the expelled exhaust. The exhaust ports are simply made short and free-flowing, and dump into an exhaust plenum that directs to a low restriction outlet pipe. Since the exhaust is comparatively cool, and since the exhaust flow velocity is low and contains minimal kinetic energy, it is expected the IPC engine will need little or no muffling.












Fig32 – Bulkhead between cyls 1&2 and 3&4 are vented to reduce crankcase turbulence in this 4-stroke IPC engine. This engine runs smoothly without counterweights or countershafts, allowing the use of unusually lightweight nodular iron crankshafts.


Combustion chamber turbulence friction reduction: Turbulence is preferably generated only from time direct fuel injection commences until the instant combustion extinguishes, to optimally pre-mix fuel and air and promote rapid combustion. Turbulence generated significantly prior to fuel injection, such as conventional tumble or swirl configurations found in many engines today, can be considered wasted vortex energy. Turbulence generated after the combustion process extinguishes is also wasted vortex energy. The IPC engine’s construction targets this goal as close as possible.

Crankcase turbulence friction reduction: Block bulkhead ports between pistons 1&2 and pistons 3&4 relieve an air pressure transient in crankcase caused when neighboring pistons travel in opposite directions. A crankcase vacuum pump may be employed to reduce crankcase windage friction.



Unfinished topics to be included in this draft



With exception of the intermediate and simplified versions of the 2-stroke, the IPC engine is an “interference engine” by design. The shown timing belt configuration should be replaced in the interference-type engines by a timing chain, since timing chains typically need no replacement during the service life of an engine.


An overpressure bypass valve, as seen in early IPC combustion chamber images, has been omitted in favor of an IPC combustion chamber fortified to withstand occasional stoichiometric combustion events, much the way ordinary Otto engine combustion chambers are fortified to sustain occasional detonation events.


A starter motor and alternator are not part of the presented IPC engine construction, as the target application is a flywheel-electric hybrid motor vehicle which multitasks the traction drive electric motor for traction drive, engine starting, and also conventional battery charging. This application uses an electrically-interfaced 500kJ/25kg carbon flywheel module for short-term traction energy storage, with flywheel energy efficiently managed through GPS terrain maps and driving history.




Limited Description of the 2-Stroke Insulated Pulse Engine

A conceptual 3.2 liter in-line 4-cylinder 2-stroke IPC engine is presented here in limited detail, since the model is now obsolete and being updated. Asample image is located at the top of this web page, and more image samples are presented below. The 3.2 liter 2-stroke is expected to produce roughly 50 horsepower at 4000 RPM, and exhaust emissions are expected to be comparable to the 4-stroke IPC engine, combusting cleanly with little or no need for emissions controls.




Introductory Image of 2-Stroke Shortblock Shown below in Fig 37 are eight rotary drum valves (green) which act as both a maintenance-free reed valve and also as a silent replacement for a roots type blower. The shutter drum valve has a uniform clearance of 0.5mm between the rotor and the block and cylinder to eliminate friction and wear, as there is no need for perfect sealing. The shutter drum valve, when shuttered closed across the intake port as the piston rises, need only apply significantly more resistance to flow than the opened exhaust valves generate, not perfect resistance to flow. As shown, each of the two shutter drum valve assembly prototypes contain six sealed ball bearings. This will be cost-reduced to three bearings per assembly when eventually stress modeled.

Prior to the shutter drum valve closing off the intake ports, the shutter drum valve scoops air ahead of it, providing a calibrated volume of positive-pressure fresh air which will assist filling the lower 1/3 of the combustion chamber as the piston falls toward BDC. The combustion chamber begins pulling a vacuum when the piston falls to 67mm ATC, as that is when the combustion reaction drops to 1 ATM pressure. When the piston reaches BDC, the top 2/3 of the combustion chamber contains exhaust, and the bottom 1/3 contains fresh air. As the piston rises, it pushes the top 2/3 combusted gas out the exhaust valves, sealing in the bottom 1/3 of fresh air for the next combustion reaction.

The shutter drum valve at the intake port differs from a roots type blower in that the shutter drum valve works to displace gasses only during the portion of the engine cycle in which the cylinder ports are exposed to the combustion chamber and able to flow, thus eliminating roots-type blower losses related to pumping against plenum pressure. Each shutter drum valve may be accompanied by a supplemental rotor, much as a roots blower uses two rotors to displace gasses, if more control of transfer volume is sought. Due to the lack of head pressure, the supplemental rotor need not contact the shutter drum valve, and therefore will be maintenance free, but should be constructed to assure the small clearance between rotors operates at tolerable leakage levels. In the intermediate version of the 2-stroke IPC engine an additional shutter drum valve applied to the exhaust port may be used to draw combusted gasses from the combustion chamber. The shutter drum valve differs from a reed valve in that it does not cyclically flex during operation and therefore requires no maintenance during the lifetime of the engine.

Since the IPC engine at full throttle consumes only 1/4 of the compressed oxygen each combustion event, the exhaust retains 75% unconsumed oxygen. Partial mixing of exhaust with inducted air while the piston rises from BDC will have minimal effect on the combustion reaction.


Fig 37 - Introductory image of twin-crankshaft shortblock for 2-stroke IPC engine. The study of twin crankshafts has concluded, and single crankshaft studies are now proceeding.


Fig 38 - Numerous aging images have been removed from this paper to improve clarity. This paper will present a full set of updated images when they become available.



End of draft of 18Apr2010, updated 04Dec2010


The Insulated Pulse Engine: A cold adiabatic engine concept
By Dave Schouweiler, shoe@bitstream.net, Minneapolis MN USA, updated 04Dec2010

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قديم 18-12-2010, 11:56 PM
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افتراضي رد: Brief Introduction to the InsuLated PuLse Engine


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This personal research project describes a thermally efficient concept for combusting fuel in an internal combustion engine. It explores adiabatic engines, a usually dormant science that was last active for several years after the 1979 oil crisis. This concept is not like published adiabatic engines which expel superheated combustion gasses into an exhaust duct for the post-processing of energy. Instead, this "cold adiabatic engine” concept allows combusted gasses to adiabatically expand and cool before exiting the combustion chamber.

This is a work in progress. This draft is a conceptual paper, not a technical paper, as it contains intuitive approximations and primitive constructions which require refinement. The images exist as props to showcase the concept, not to suggest a best design practice. I’m not an engine designer and I don't claim this concept works. I only seek to learn why this concept may or may not work, perhaps find the flaw or fallacy which is understood by industry, and publish the results in this paper. This engine concept can be fabricated using century-old technology, and similar concepts have certainly been studied and dismissed, but the findings are not readily available. Technical critique and information on similar experiments is welcome.

Brief Introduction to the Insulated Pulse Engine

The "insulated pulse-combustion engine", abbreviated "insulated pulse engine" or "IPC engine", is a low volumetric efficiency internal combustion engine concept which combusts fuel at high thermal efficiency. The IPC engine sequence is thermodynamically quite simple, applying comparatively pure forms of: 1) isentropic (adiabatic) compression, 2) isochoric (pulsed) heat addition, 3) isentropic (adiabatic) expansion, and 4) isobaric heat rejection.

The P-V diagram of an ideal Otto cycle (not shown, but readily searchable) indicates the Otto (gasoline) engine also has a purely isochoric heat addition process, however it is well recognized the Otto engine more accurately incorporates a two-stage heat addition process which starts out isochoric and transitions to isobaric, with the isobaric segment significantly increasing volumetric efficiency while significantly decreasing thermal efficiency. The IPC engine concept principally differs from the Otto engine in that it:
1. Optimizes isentropic compression (much the way a Diesel engine does),
2. Eliminates the isobaric component of heat addition (much the way HCCI engine prototypes do),
3. Significantly extends isentropic expansion (much the way an Atkinson engine does), and in doing so,
4. Enables practical thermal insulation of the combustion chamber (much the way adiabatic engine prototypes did, except the IPC engine is able to use cheap durable insulators).


The combustion chamber of the IPC engine isselectively insulated using two economical Fe60Ni40 alloy inserts to minimize heat energy loss to a cooling system. Combustion initiates and is consumed rapidly near top dead center (TDC), permitting adiabatic cooling of combustion chamber gasses through the entire expansion cycle. The expansion cycle is extended beyond convention to extract additional heat energy from the combusted gasses, further reducing average combustion chamber temperatures to minimize stress on the thermal insulators, eliminating the need for a cooling system and resulting in an exhaust that is comparatively cool and pressureless.


Conventional emissions control devices don't work with low temperature exhaust gasses, so the IPC engine stratifies fuel to locally combust in a region of the combustion chamber specifically shaped to support efficient, clean combustion. Fuel stratification additionally permits throttling the homogenous fuel-air equivalence ratio within the highly reactive range of 0.40-0.80 to assure a rapid, complete combustion reaction while retaining a useful power band. A fuel-air equivalence ratio below 1.00 represents the deviation of a stoichiometric ratio toward fuel-lean.

Due to stratification constraints, the crankshaft RPM limit is defined by the combustion reaction rate of the selected fuel. Operating above the RPM limit of the selected fuel promotes incomplete combustion. When operated at or below the RPM limit, the IPC engine combusts cleanly with minimal need for emissions controls. For investigative purposes, gasoline is assumed the fuel used in the IPC engine, but diesel, propane, ethanol, ammonia, or most any conventional motor fuel is applicable to the IPC engine.


This website introduced a hypothetical 3.2 liter in-line 4-cylinder 4-stroke IPC engine with twin counter-rotating crankshafts on 04Sep2009 that generated an expected 50 HP at 4000 RPM. This website partly introduced a hypothetical 3.2 liter in-line 4-cylinder 2-stroke IPC engine on 08Apr2010 that generatedsimilar power. Much was learned about engine structures as design boundaries were stretched in these early models. Current activity is to reintroduce these engines in a more compact and practical "single crankshaft" form. For investigative purposes, the cylinder bore diameter of the IPC engine in this paper is 100mm and the piston stroke is 100mm.

When compared with naturally-aspirated Otto and Diesel engines at full throttle, a similarly displaced IPC engine at full throttle consumes roughly an eighth of the fuel each combustion event. This is based on the observation that HCCI prototype engines at full throttle consume a fourth of the fuel that Otto or Diesel engines of similar displacement consume at full throttle, and only half the piston stroke of the IPC engine is used during the compression cycle. The IPC engine is expected to have almost twice the peak fuel efficiency of Otto and Diesel engines, and will therefore generate roughly a fourth of the horsepower of similarly displaced Otto and Diesel engines at full throttle and similar RPM. Due to the increased thermal dwell of each combustion event, the IPC engine at reduced RPM is expected to have greater than twice the fuel efficiency of Otto and Diesel engines at reduced RPM.

The cylinder displacement requirements of an IPC engine are roughly four times that of Otto and Diesel engines at equivalent horsepower and RPM, but the cost, weight, and space requirements of the IPC engine assembly remain comparable due to a reduction in need for cooling, muffling, and emissions control components. Since mechanical friction is a variable which correlates more closely to generated horsepower than to displacement, and since the IPC engine is constructed using methods which emphasize reduction of mechanical friction, windage friction, and port pumping losses, friction generated within the IPC engine is comparable to friction generated within equivalently powered Otto and Diesel engines.
Fig10 – Introductory image of a deluxe version of the 3.2 liter, 50 horsepower, 2-stroke IPC engine concept, the deluxe version containing intake ports in the cylinder bore and exhaust poppet valves in the head. Twin crankshafts, as shown, reduce piston skirt friction and piston noise, but are found to undesirably increase engine cost, size, and weight. A single crankshaft is now the preferred direction for IPC engine modeling. The next IPC engine to be presented here will be an intermediate version of the 2-stroke which, because it requires no poppet valves or camshafts, is expected to be a simple, lightweight, low friction, low maintenance, and very low emissions engine. Both a tuned exhaust (shown in brown) and an untuned exhaust have been modeled for investigative purposes. The next IPC engine models will employ the untuned, free-flow exhaust, since a tuned exhaust adds cost but provides negligible performance benefit in an engine with a cool, pressureless exhaust.



Volumetric Efficiency and Thermal Efficiency

There is an ongoing effort to improve fuel mileage in motor vehicles. In the last half century, fuel mileage improvements from internal combustion engines have most often resulted from increased volumetric efficiency (i.e.: increased peak horsepower per unit volume of cylinder displacement), not increased thermal efficiency. Fuel mileage gains have come from a deliberate matching of small displacement engine to large vehicle such that an engine issimply tasked to operate within a more thermally efficient segment of its operating range.


Higher volumetric efficiency in modern engines does not indicate improved thermal efficiency in an engine. For example, an older 80 horsepower 2 liter engine and a modern 160 horsepower 2 liter engine will likely provide about the same fuel mileage in a particular small car application.

Small displacement engines with high volumetric efficiency operate at higher combustion chamber temperature and pressure and higher RPM than do similarly tasked large displacement engines, reducing combustion chamber surface area and reducing exposure time in which each combustion event can lose heat energy to a cooling system. These conditions keep a 160 horsepower 2 liter engine within a more thermally efficient segment of its operating range when matched to a large vehicle, leading to better fuel mileage than achievable with a 160 horsepower 4 liter engine in the same large vehicle.

Cooling System Efficiency Losses

Internal combustion engines incorporate a cooling system to quickly remove heat energy absorbed by combustion chamber metals after each combustion event. This removal is necessary, since chamber metals would otherwise attain the average temperature of the combustion chamber gasses, a temperature too hot in Otto and Diesel engines for sustainable engine operation. Heat energy conducted through the combustion chamber metal into the cooling system represents a significant reduction in the thermal efficiency of an engine, particularly at low RPM when the dwell time for each combustion event is longest.

Following the oil crisis of 1979, internal combustion engine manufacturers around the world began developing “adiabatic engine” prototypes which contained thermally insulated ceramic combustion chambers in an attempt to improve engine thermal efficiency without sacrificing volumetric efficiency. Thermally insulating the combustion chamber reduced, and sometimes eliminated, the need for a cooling system, thus retaining a larger fraction of combustion heat energy for mechanical work output. In order to retain volumetric efficiency, these adiabatic engines were designed to combust with a conventional low heat release rate. This low heat release rate superheated the combustion chamber gasses before expelling them into the exhaust duct for the purpose of energy recovery through turbocompounding and other post-processing methods.

Experimental results on three published ceramic adiabatic engine projects can be reviewed in SAE technical papers 810070 (1981), 820431 (1982), and 840428 (1984), with abstracts viewable at the SAE.org website and where the papers may be downloaded. Adiabatic engines of the 1980s operated under the most brutal conditions. Adiabatic engines provided improved fuel efficiency, but could not be made sufficiently reliable for commercial application.

The use of a ceramic material, or the use of any thermally insulating material, to insulate combustion chambers of internal combustion engines for the primary purpose of improving fuel mileage in vehicles has found minimal research interest in the industry since the conclusion of these experiments.

The ceramic adiabatic engine experiments being developed around the world in the early 1980s were much more secretive than the prolific adiabatic engine experiments of race car builder Smokey Yunick. Smokey's many different adiabatic engine prototypes, which became staples of popular automotive magazines from 1979 through 1984, similarly blended the stressful combination of high volumetric efficiency with increased thermal efficiency, and then added some racing magic. Material longevity may have been one reason his adiabatic engines failed to achieve commercial applicability, and streetability may have been another. The magazine articles didn't provide in-depth or follow-up analyses, but watchful engine manufacturers certainly got all the necessary detail.

Exhaust System Efficiency Losses

In both Otto and Diesel engines, and in the adiabatic engine experiments described above, combustion is engineered to progress gradually, beginning near TDC and continuing well into the expansion cycle. This low heat release rate allows a lot of fuel to gradually burn without exceeding the pressure limits of the combustion chamber, providing high volumetric efficiency and low thermal efficiency. Volumetric efficiency is high because the piston experiences high levels of combustion pressure through a significant portion of the expansion cycle. Thermal efficiency is low because the late burning fuel cannot adiabatically expand as many times as the early burning fuel. Late burning fuel causes large amounts of fuel energy to be lost to the exhaust in the form of heat and pressure.

Unfortunately, the quantity of fuel Otto and Diesel engines must consume to generate high levels of horsepower cannot be combusted entirely at TDC without exceeding the pressure limits of the combustion chamber, so Otto and Diesel engines intentionally reduce the burn rate to achieve high volumetric efficiency. As applied in the adiabatic engine experiments of the 1980s, this longer burn duration exposed ceramic combustion chamber surfaces to more heat energy, raising temperature gradients within the body of the ceramic. The lower heat release rate may have set up thermal gradient stresses within the ceramic which contributed to reduced ceramic durability. By contrast, HCCI engine prototypes in research laboratories today combust all fuel near TDC and none during the expansion cycle, and Atkinson engines extend the expansion cycle until useable combustion pressure is mechanically consumed. These latter two engines release less heat energy to the exhaust than do equivalently powered Otto, Diesel, and adiabatic engines.




The Thermal Efficiencies of Insulation, Combustion, Expansion, and Friction
Thermal efficiency in an internal combustion engine is comprised of the following four core efficiencies, the first three of which determine the "adiabatic efficiency" of an engine:
1. High "insulation efficiency" minimizes loss of combustion energy to a cooling system in the form of heat, and is driven by the thermal conductivity of the combustion chamber. If maximizing thermal efficiency is the primary goal, and if excessive heat is lost to a cooling system, the insulation efficiency must be improved. If improved insulation efficiency causes the combustion chamber material to overheat and fail, the average temperature of combustion chamber gasses through a full engine cycle must be reduced.
2. High "combustion efficiency" minimizes loss of combustion energy to the exhaust duct in the form of elevated exhaust temperature, and is driven by compression ratio, ignition timing, and combustion duration. If maximizing thermal efficiency is the primary goal, and if the temperature of combustion chamber gasses is excessive at the end of the expansion cycle, the combustion efficiency must be improved.

3. High "expansion efficiency" minimizes loss of combustion energy to the exhaust duct in the form of elevated exhaust pressure, and is driven by the expansion ratio. If maximizing thermal efficiency is the primary goal, and if the pressure of combustion chamber gasses is excessive at the end of the expansion cycle, the expansion efficiency must be improved.

4. High "friction reduction efficiency" minimizes loss of combustion energy to mechanical component friction and to air pumping friction within the engine.


Adiabatic efficiency of Otto and Diesel engines
Insulation efficiency is low in Otto and Diesel engines due to the high average temperature of combustion chamber gasses. The high average temperature of gasses necessitates the active cooling of combustion chamber materials to keep them at reliable operating temperatures, resulting in significant combustion energy loss to the cooling system. Combustion efficiency is low in Otto and Diesel engines because high volumetric efficiency does not allow full combustion at TDC without developing excessive cylinder pressure, creating the need for a combustion process with a gradual "low heat release rate", in which the latter combusting fuel expands at lower adiabatic efficiency than the earlier combusting fuel, the latter combusting fuel providing significant heat energy loss to the exhaust duct. Expansion efficiency is low in Otto and Diesel engines because the compression process and expansion process are conveniently of equal stroke length, resulting in significant pressure energy being released to the exhaust duct before it can perform work on the piston. It should be noted that the compression cycle and expansion cycle are independent functions and will seldom be of equal length in an engine optimized for high thermal efficiency.

Adiabatic efficiency of the ceramic adiabatic prototype engine
Insulation efficiency was high in the adiabatic engine experiments of the early 1980s, but combustion efficiency and expansion efficiency were both low. These “adiabatic engines” were, in effect, one-third adiabatic, not fully adiabatic. Thermal efficiency was high because the thermally insulating ceramic combustion chamber material prevented significant heat energy loss to a cooling system. Combustion efficiency was low because the fuel in these experiments burned with a low heat release rate in order to maintain conventional high volumetric efficiencies, with the latter burning fuel generating significant heat energy loss to the exhaust duct. Only the small portion of fuel burning near TDC combusted at high adiabatic efficiency. Expansion efficiency was low because significant pressure energy remained in the combustion chamber when the exhaust cycle began, with only a fraction of this energy recovered through post-processing. The combined result was a brutally hot expansion and exhaust process which provided some improvement in thermal efficiency over Otto and Diesel engines, but which excessively stressed the ceramic thermal insulators, rendering them unsuitable for commercial application.

Adiabatic efficiency of the HCCI prototype engine
Combustion efficiency is high in HCCI engines being researched around the world today, but expansion efficiency and insulation efficiency are both low. Combustion efficiency is high because the entire combustion reaction occurs at a “high heat release rate” near TDC, letting all fuel perform work on the piston from the start of the expansion cycle. Expansion efficiency is low because useable pressure remains in the combustion chamber when the exhaust cycle begins, resulting in the loss of significant pressure energy to the exhaust duct before it can perform work on the piston. Insulation efficiency is low in the HCCI engine, since an active cooling system is required to keep combustion chamber materials at a reliable operating temperature. The HCCI engine can also be considered a "one-third adiabatic" engine.

Adiabatic efficiency of the Atkinson engine
Expansion efficiency is high in Atkinson engines being produced today, but combustion efficiency and insulation efficiency are both low. Insulation efficiency is low in Atkinson engines, since an active cooling system is required to keep combustion chamber materials at reliable operating temperatures. Combustion efficiency is low, because fuel must combust at a thermally inefficient "low heat release rate" through a significant portion of the expansion cycle. Expansion efficiency is high, in that the expansion cycle is extended in stroke length beyond that of the compression cycle, allowing extraction of all useable pressure energy from the combustion chamber. Thisdefines the Atkinson engine as yet another type of “one-third adiabatic” engine.

Adiabatic efficiency of the Insulated Pulse conceptual engine
Insulation efficiency, combustion efficiency, and expansion efficiency are all high in the IPC engine, and the constructions described below will provide a notable increase in fuel efficiency over adiabatic, HCCI, and Atkinson engines while combusting cleanly, without need for pollution controls. The IPC engine is a true adiabatic engine construction, but to prevent confusion with established naming practice, the IPC engine is probably best called a “cold adiabatic engine”, since it transfers minimal heat to a cooling system and expels minimal heat energy and minimal pressure energy out the exhaust duct. It should be noted that, while the adiabatic efficiency is high in the IPC engine, "friction reduction efficiency" must be included to determine thermal efficiency.

Industrial Trends

The current emphasis of industry is to follow the path of high volumetric efficiency to improve fuel mileage in motor vehicles, however fuel mileage gains may become tougher to find assmall engines more routinely populate large vehicles. Atkinson engines, which achieve improved thermal efficiency through reduced volumetric efficiency, are found in some of today’s most fuel efficient cars. HCCI engine development programs, now popular in laboratories around the world, seek high thermal efficiency using a process which has low volumetric efficiency. These two enginessuggest the possibility that low volumetric efficiency may eventually lead the way toward significant improvements in engine thermal efficiency and fuel mileage.





Fig12 – Gas ported piston with low tension compression rings and reduced skirt contact. Presently, gas ported pistons are only applicable to racing engines, since any fuel entering the gas ports will remain uncombusted and be ejected as a pollutant. Since the stratified combustion chamber of the IPC engine does not allow fuel to enter the ports, gas ported pistons are not a pollution issue in an IPC engine, nor are the ports subject to clogging.


Exhaust Emissions

Exhaust emission concerns in the insulated pulse engine fall into four simplified categories:
1. Hydrocarbon (HC) exhaust emissions, representing fuel that is not combusted, are formed when fuel is in proximity of chilled combustion chamber crevicessuch as are found near the head gasket, upper piston ring, and intake valve seat.

2. Soot emissions, also known as particulate matter (PM) emissions, representing fuel that is 1/3 combusted, are formed when fuel is direct injected into the dense flame kernel of a compression ignition engine which has already consumed all adjacent oxygen.

3. Carbon monoxide (CO) emissions, representing fuel that is 2/3 combusted, are formed when fuel is combusted near chilled surfaces within the combustion chamber.

4. Oxides of nitrogen (NOx) emissions are generated when heat energy becomes unnecessarily high in the combustion chamber and the very stable 3-bond nitrogen molecule breaks apart.

The cause of exhaust pollution in internal combustion engines is complex but well understood, as are clean combustion methods which prevent pollution, and as are exhaust processing methods which remove pollution.

Constructions which promote clean combustion have been extensively adopted by the IPC engine, since the cool temperature of the IPC engine's exhaust renders many popular emissions control devices ineffective, as many depend on significant levels of exhaust heat to function. Combustion in the IPC engine issufficiently unique that some form of emissions control will likely be required, but emissions levelsshould be sufficiently low that incorporation of the needed controls will not significantly affect cost or thermal efficiency.

Exhaust emissions in the IPC engine concept can be characterized more clearly once the initial energy calculations for an IPC engine cycle are completed. Before this will happen, an acceptably mature IPC engine assembly will be modeled in CAD, followed by presentation of the energy equations and initial calculations, using the CAD assembly model to base the calculations. At present, there exists little formal math to support the claims and curiosities of this paper. This conceptual paper will become a technical paper once the initial energy calculations are posted to this website. The focus will then turn to formatting this website for presentation and continued concept characterization.

There may eventually be opportunity to run engine simulation software on the IPC engine concept which will move the energy, emissions, and stress calculations closer to simulating a physical engine, thereby providing an industry-accepted baseline which effectively proves or disproves the engine concept on paper. Whether proved or disproved, if this project gets to the point of running simulations, the simulation results will be published. If proven and published, expanded research interest might commence. If disproven and published, future investigation into this conceptual construction would be resolved in short order. To be sure, there is a lot to study, calculate, and learn from each stage of this development process.

Basic Description of the Insulated Pulse Engine

The insulated pulse engine is an ordinary reciprocating piston internal combustion engine which applies unthrottled air induction, direct fuel injection, spark ignition, and the following three unconventional functions, to achieve high thermal efficiency:
1. Rapid "pulse" combustion (like an HCCI engine).
2. Thermally insulated combustion chamber (like an adiabatic engine).
3. Extended expansion cycle (like an Atkinson engine).

These three unconventional functions combine to create an engine with both high thermal efficiency and low volumetric efficiency. Mechanical friction and windage friction take on greater significance in engines with reduced volumetric efficiency. The insulated pulse engine carefully manages friction reduction and cost to optimize thermal efficiency. Friction reducing methods to consider include:
1. Twin counter-rotating crankshafts eliminate piston side thrust friction.
2. With single crankshaft, a longer connecting rod reduces piston side thrust friction.
3. Anti-friction material applied to piston skirt reduces piston side thrust friction.

4. Reducing excessive piston skirt contact area reduces viscous friction.
5. Gas ported low-tension piston rings reduce piston sliding friction.
6. Minimize port flow volume and port flow resistance, avoid throttled induction.
7. Minimize crankcase windage using crankcase vacuum and strategic bulkhead vents.
8. Turbulence should mix fuel with air efficiently, not excessively.
9. Rolling contact bearings, where possible, consume less energy than friction bearings.

The resulting engine requires only an active oil cooler of ordinary capacity to support all cooling needs, does not require a muffler to function quietly, and exhaust gasses can be made sufficiently cool that the exhaust manifold may be molded of plastic.

Twin counter-rotating crankshafts have been investigated here as a potentially cost-effective method to reduce friction within the IPC engine. Twin crankshafts hold additional value in noise reduction. Though the IPC engine does not generate unusually high combustion chamber pressures, it does generate an unconventionally rapid rate of change in pressure near TDC, and this may promote an undesirable "slap" sound if side thrust is applied to the piston by the connecting rod due to application of a single crankshaft. The present modeling constructions suggest twin crankshafts may unacceptably add to the height, width, weight, and cost of the engine. Development has now shifted toward a more practical single crankshaft construction using a short steel connecting rod in combination with an aluminum block. Whether or not any of the friction reducing suggestions shown above are applied to the IPC engine is not actually material to the core engine concept. It is only necessary to observe that friction reduction will improve thermal efficiency, and cost-benefit studies will determine which suggestions merit incorporation.




Fig13 – The study of one form of a piston and rod assembly for a dual crankshaft construction indicates engine size and weight may be undesirably increased. Current direction targets a single crankshaft design.


1) Rapid “Pulse” Combustion

In the IPC engine, combustion initiates near TDC and is rapidly consumed near TDC, providing combustion with low volumetric efficiency and high thermal efficiency. The volumetric efficiency is low because a comparatively small amount of fuel will generate sufficient temperature and pressure near TDC to reach the limits which do not form NOx exhaust pollutants. Thermal efficiency is high because all of the combusted gasses adiabatically cool through the entire expansion cycle, greatly reducing the percentage of heat energy lost out the exhaust and lowering the average temperature of the combustion chamber. The ordinary methodsselected to achieve this high heat release rate are:
1. High compression ratio
2. Combustion chamber shaped to fully support efficient combustion
3. Fuel-air charge turbulently mixed prior to ignition
4. Combustion chamber turbulence present at time of ignition
5. Additional combustion chamber turbulence generated by combustion reaction
6. Fuel-lean equivalence ratio optimized for rapid, complete reaction
7. Spark ignition precisely controls combustion envelope
Complete combustion at TDC in the IPC engine does not generate destructive pressure, as there is an insufficient quantity of fuel in the combustion chamber during each combustion event to generate excessive pressure. Pressure and temperature limits in the IPC engine’s combustion chamber are not driven by structural limits, but are driven by the need to prevent the formation of NOx emissions during combustion. If temperature and pressure in the combustion chamber climb sufficiently high that the very stable 3-bond nitrogen molecule breaks apart and forms NOx emissions, then temperature and pressure must be readjusted below NOx-producing levels, since the IPC engine is intended to combust cleanly without pollution controls.


Engine misfire may occasionally cause an anomalousstoichiometric fuel-air mixture to combust at detonation pressures in the chamber. The IPC engine, like conventional engines, is constructed to handle this type of misfire condition.




Fig14 – Piston shown 9mm BTC, just as combustion chamber becomes fully insulating.

2) Thermally Insulated Combustion Chamber

The IPC engine thermally insulates the combustion chamber completely when the piston is within 9mm of TDC, and partly insulates when the piston is further than 9mm from TDC. Three reasons for insulating are to 1) increase thermal efficiency by minimizing heat energy loss to the cooling system during the hottest portion of the compression and expansion cycles, 2) to burn cleanly at TDC by assuring critical combustion chamber surfaces flash to higher temperatures during compression and combustion to prevent the formation of CO exhaust emissions, and 3) to bring the combustion chamber up to operating temperature as fast as possible to minimize HC and CO exhaust pollutants at cold engine start-up.

The preferred thermal insulating material in the IPC engine is an iron or steel alloy containing 40% nickel of 3mm thickness, with thermal conductivity of 10 W/m K at 200 degrees C. As a comparison, the thermal conductivity of cast A356-T6 aluminum is 130 W/m K at 200 degrees C with typical thermal gradient distance of 10mm between combustion chamber and cooling system, and compacted gray iron is 40 W/m K with typical gradient distance of 5mm.A ceramic popular in the adiabatic engine prototypes of the 1980s, with thermal conductivity of 2 W/m K, is reserved as a preferred insulator in a future state of development of the IPC engine concept. More research is needed before this ceramic, known as partially stabilized zirconia (PSZ), can be proven reliable. Powdered metal-ceramic composites and other combustion resistant thermally insulating materials may also find future value as a thermal insulator in the IPC engine.Discrete ceramic thermal insulators may also improve CO exhaust emissions, due to more significant heating of the combustion chamber surfaces during compression and combustion, but turbulence during compression and combustion is expected to heat thermally insulating nickel steel combustion chamber surfacessufficiently to minimize CO exhaust emissions. Selective application of commercially available ceramic film coatings to the nickel steel combustion chamber may alternately minimize CO emissions, if needed.

As indicated, there are thermally insulating materials which insulate better than the selected 40% nickel steel alloy, but ideal insulators will not substantially improve engine thermal efficiency. The selected nickel steel will perform nearly as well as an ideal thermal insulator at high engine RPM, and will only become significantly less efficient than ideal insulators at low engine RPM, when the heat energy of each combustion event has more time to be absorbed by combustion chamber material. Even at low RPM, the nickel steel combustion chamber remainssignificantly more thermally efficient than Otto and Diesel combustion chambers. The rate of fuel consumption is low at low RPM, and therefore application of non-ideal insulators has reduced significance.

3) Extended Expansion Cycle

The IPC engine incorporates an extended expansion cycle, much like an Atkinson engine, to let combustion energy perform additional motive work before being discharged to the exhaust. The extended expansion cycle further reduces average combustion chamber temperature, bringing the average combustion chamber temperature down to the level where a cooling system is not required at all, except perhaps when running at full throttle in hot ambient conditions. When cooling is required, excess heat is readily removed via an external oil cooling system of ordinary capacity.An expansion ratio value isselected which will assure expansion energy gains constructively exceed friction force losses through the entire expansion cycle. Fuel pricing may apply a market force which influences the final specification of the expansion ratio. Otto and Diesel engines have evolved such that the compression and expansion cycles are approximately matched in stroke length. The compression cycle and the expansion cycle are each driven by significantly different physical parameters and mathematical equations, and their lengths will seldom coincide if maximized thermal efficiency is a primary goal. In the 4-stroke IPC engine, the compression cycle is roughly half of a piston stroke and the the expansion cycle is roughly a full piston stroke. In the 2-stroke IPC engine the compression cycle is roughly one third of a piston stroke and the expansion ratio is roughly two-thirds of a piston stroke.

Fuel-Stratified Combustion Chamber

Two critical issues exist with the combustion process described for the IPC engine: First, complete full-throttle combustion at TDC with a stoichiometric mix of fuel and air generates destructive pressure levels. With gasoline selected as the fuel, calculations indicate that a fuel-lean equivalence ratio of no more than about 0.25 is required to prevent excessive cylinder pressure when all fuel combusts at TDC, but full-throttle equivalence ratios in this low range combust incompletely and generate significant CO exhaust pollutants, with part-throttle equivalence ratios becoming non-combustive, as demonstrated in HCCI engine prototypes. Second, with homogenously mixed combustion reactions, there exist chilled locations in the combustion chamber which don’t support efficient combustion, yet which contain fuel and air. Examples of these locations include the clearance volume between the O.D. of the piston and I.D. of the cylinder bore above the sealing rings, the surface of the head gasket exposed to the combustion chamber, and the junction adjacent to the intake valve and seat. HC pollution is created in these locations of a homogenously inducted combustion chamber.

A stratified combustion chamber can resolve both the destructive pressure and the incomplete combustion issues. By splitting the combustion chamber into two separate regions just prior to direct fuel injection, one region can be designed to contain only air, while the other generate a turbulent mix of fuel and air, permitting clean and fast combusting fuel-air equivalence ratios in the range of 0.40-0.80 while segregating "chilly" chamber featuresinto the air-only region of the combustion chamber. The fuel-air region can be optimally shaped to fully support combustion, and a transfer passage between the two regions can be designed to direct turbulent kinetic energy prior to ignition and then support efficient expansion of the combusting reaction.

The stratified combustion chamber for the IPC engine initially forms when the piston is within 12mm of TDC and becomes shaped for clean fast combustion only when the piston is within 0.5mm of TDC. Spark ignition is required to assure combustion occurs precisely within this positional constraint. The rate of the combustion reaction is driven, in part, by the selected fuel, the compression ratio, the fuel-air equivalence ratio, chamber turbulence, and engine RPM, and will require a specified length of time to burn completely and cleanly. This reaction time defines an engine RPM maximum which, if exceeded, will result in pollution emissions. The IPC engine operates with greatest thermal efficiency at or just below this RPM maximum, and it retains practical levels of thermal efficiency at lower RPM. A maximum RPM value of 4000 has arbitrarily been assigned to the IPC engine for investagive purposes.




Fig15 - Thermally insulating piston cap, made from investment cast 40% nickel steel.



Fig16 – Back view of piston cap and head dish showing insert-casting retention features.



Fig17 – Thermally insulating 40% nickel steel head dish with thermal conductivity of 10 W/m K. Poppet valves are omitted from some versions of the 2-stroke IPC engine, significantly simplifying this casting.


Detailed Description of the 4-stroke Insulated Pulse Engine

Insulated Combustion Chamber
The most significant feature of the IPC engine is the thermally insulated combustion chamber: The piston contains an insulating cap, and the cylinder head contains an insulating dish. This is the full extent of the insulation. The unique size and shape of the stratified combustion chamber has forced a reduction in the valve head diameter, requiring a multi-valve arrangement to retain low-restriction intake and exhaust flow. The shown configuration of eight valves per cylinder is primarily a research tool, intended to assure sufficient flow for early experimentation and characterization. The combustion chamber will later be optimized to reduce cost.

These two insulating components will be investment cast out of a nickel iron or nickel steel alloy chosen for low thermal conductivity, high temperature stability, and valve seat wear resistance. For valve seat wear resistance, there issome carbon added to permit localized induction hardening. One of these insulators is inserted into the die cast mold of an aluminum piston to keep reciprocating mass low, the other is inserted into the mold of a cast aluminum cylinder head to keep engine mass low. The head insulator combines the duties of valve seat, spark/injector mount, and head gasket sealingsurface.

The valves installed into this head assembly are made of an inexpensive stainlesssteel or nickel steel alloy chosen for comparatively low thermal conductivity and for tribological compatibility with the nickel steel valve seats. The cylinder bore and piston rings are cast of conventional engine materials to assure good lubricity and long life at minimal cost.

While a steel containing 40% nickel, processed to produce a microstructure with thermal conductivity of approximately 10.0 W/m K at 200 degrees C, is the preferred insulating material in the IPC engine prototype, it may not be the best in future applications. Discrete ceramic components may provide improved insulating performance, but are brittle and will require a significant development effort to reliably incorporate into this engine.

The preferred discrete ceramic used in the adiabatic engine experiments of the early 1980s is called partially stabilized zirconia (PSZ). SAE Technical Papers 820429 (1982) and 830318 (1983), with abstracts viewable at the SAE.org website and where the papers may be downloaded, discuss internal combustion engine uses for discrete PSZ ceramic components. The thermal conductivity of two preferred PSZ ceramic compositions is approximately 2.0 W/m K.

PSZ ceramic was not sufficiently durable in the adiabatic engine experiments to become commercially applicable, though it performed remarkably well considering the severity of testing. It is expected PSZ will perform quite reliably at the lower average combustion chamber temperatures and milder thermal gradients within the IPC engine, but it must be incorporated in a manner which applies minimal tensile loading, preferring compressive loading where loads must exist.

Discrete ceramic insulators will not likely improve thermal efficiency in the IPC engine by more than a few percent over the selected nickel steel, but the lower thermal conductivity of discrete ceramic insulators may improve thermal efficiency of the engine at lower RPMs, such that the increased dwell of each combustion event at lower RPMs does not result in significantly increased heat energy absorption into the combustion chamber material. Discrete ceramic insulators may also improve CO exhaust emissions, but it is expected that turbulence during compression and combustion will heat critical nickel steel combustion chamber surfaces sufficiently to prevent CO exhaust emissions.

A queued task is to determine whether the thermally insulating microstructure of the selected 40% nickel alloy is stable in the stressful operating environment of the IPC engine, or whether the selected alloy reverts to a more thermally conductive microstructure. Actual choice of insulating alloy is not a core issue, and will be determined when insulating material becomes the research focus for the IPC engine concept.

Because the cylinder of the IPC engine is made of conventional materials which are thermally conductive, the combustion chamber will only be fully insulating when the piston is within 9mm of TDC. With the brief combustion reaction near TDC, combustion chamber temperatures drop considerably by the time the cast iron cylinder bore issignificantly exposed to combustion chamber gasses, minimizing heat energy loss.

The combustion chamber predominantly insulates when the piston is within 9mm of TDC. The combustion chamber partially insulates when the piston is farther than 9mm from TDC. As the piston travels from TDC toward BDC, and while the piston remains closer than 9mm to TDC, the heat generated by the combustion reaction is almost entirely dedicated to applying force to the crankshaft, finding minimal opportunity to route heat energy to a cooling system.

The combustion chamber switches from predominantly insulating to partially insulating when the piston drops below 9mm from TDC, as a segment of thermally conductive cast iron cylinder bore starts to occupy a small portion of the combustion chamber'ssurface area. Combustion chamber gasses have adiabatically dropped in temperature by the time the thermally conductive cylinder bore surface becomes a significant percentage of the combustion chamber surface area, greatly reducing heat energy absorption into the cast iron cylinder.

The thermally insulating segments of the combustion chamber exist to reduce heat energy absorption, thereby preserving heat energy for mechanical work, and to assist with complete combustion to minimize pollutant emissions. The thermally conductive segments of the combustion chamber exist in order to circumvent the tribological development requirements associated with using thermally insulating materials as wear surfaces.




Oil Cooling
Since the thermally conductive cast iron cylinder bore cyclically forms a portion of the combustion chamber, it absorbs a small portion of the heat of combustion. The average cyclic temperature of the cast iron cylinder bore remains below that which requires active cooling. The thermally insulating portion of the combustion chamber slowly absorbssome of the heat of combustion and needs to transfer this heat away. The cooling method is managed by ordinary oil circulation within the engine. The oil circulation system assures all parts of the engine are lubricated as required, and all are kept at functional temperatures. Should the oil temperature climb to a designated upper limit, an external oil cooling circuit will activate. This remote cooling circuit includes a small radiator and blower fan. When wind and cold weather are present, the IPC engine is best suited to operate in an enclosure without ambient venting, to prevent overcooling.

Since the IPC engine can be operated in conditions where the oil temperature remains cool for extended periods (cold climates, short trips), the oil may become saturated with water and degrade. An oil heat exchanger can be incorporated adjacent to an exhaust duct, and exhaust gasses can temporarily be routed through the oil heat exchanger whenever oil is below a specified minimum operating temperature. Since reactive combustion energy does not contact the cylinder bore in an IPC engine, cylinder bore oiling requirements are not assevere as those in conventional engines in which a flame contacts the internal bore.










Stratified Combustion Chamber
The uniquely shaped combustion chamber of the IPC engine forms a small but significant volume between the O.D. of the piston and I.D. of the cylinder bore above the compression sealing rings. Thissmall cylindrical volume is not shaped to support efficient combustion, and will generate pollution emissions if fuel is allowed to occupy this volume. Similar inefficient volumes within the combustion chamber exist at the head gasket and valve seats.

Modern Otto cycle engines design the pistons to minimize the inefficient cylindrical volume above the compression sealing rings, and the few exhaust emissions forming in the existing small volumes are scrubbed clean by a catalytic converter. Minimizing this volume in an IPC engine requires a reduction of thermal insulation coverage, in order that the sealing rings can be located as close as possible to the compression end of the piston This can reduce thermal efficiency of the engine. Additionally, the IPC engine generates a comparatively cool exhaust when compared with Otto and Diesel engines, and conventional catalytic converters do not perform efficiently at these lower exhaust temperatures. For this reason, the IPC engine must take another approach to eliminating crevice-sourced pollutants.

The IPC engine is designed to prevent the creation of pollution in areas of the combustion chamber which don’t support efficient combustion, since it is designed to keep direct injected fuel out of these locations. The established way to keep fuel away from these locations is to operate as a Diesel cycle engine, spontaneously combusting direct injected fuel as it enters the combustion chamber, but Diesel engines intrinsically suffer from soot emissions, since fuel must be injected directly into the center of a dense flame kernel which has already consumed all adjacent oxygen. Diesel engines must remove soot pollution from the exhaust using a particulate burner, but a particulate burner does not function efficiently with the comparatively cool exhaust of the IPC engine.

The solution for preventing the creation of pollution emissions in the IPC engine is found in combustion chamber stratification. Combustion chamber stratification, in coordination with an insulated combustion chamber, pulse combustion, uniquely timed direct injection, and spark ignition, create a combustion environment which favors clean combustion and minimizes the creation of exhaust pollutants, minimizing the need for emissionscontrols.

The combustion chamber of the IPC engine isstratified only when the piston is located within 12mm of TDC. When the piston is farther than 12mm from TDC there exists only one region in the chamber. The stratified combustion chamber forms when the piston is at 12mm BTC, segregating into a perimeter squish region (called a "crevice chamber" in older images) which actively rejects fuel and a central combustion region (called a "central combustion chamber" in older images) which is optimized to mix injected fuel with air and combust cleanly. An annular transfer passage (called an "annular passage" or a "backfill passage" in older images) communicates between the two regions, transferring air toward the central combustion region as the piston rises above 12mm BTC, returning fully combusted gasses to the perimeter squish region as the piston falls to 12mm ATC. The annular transfer passage also provides a buffer at TDC which efficiently constrains the combustion reaction.

The perimeter squish region assists complete combustion: It keeps fuel away from combustion chamber features which do not efficiently support combustion. While the piston approaches TDC the perimeter squish region acts as an air pump which transfers air toward the central combustion region to turbulently mix injected fuel with air prior to ignition. Direct fuel injection begins when the piston is 8mm BTC and ends by 6mm BTC. The direct injector nozzles are aimed to inject fuel mass only into the piston pocket at the center of the central combustion region. The air pumping action actively constrains all direct injected fuel to the central combustion region, permitting selection of preferred fuel-air equivalence ratios in the range of 0.40 to 0.80 which combust most rapidly and cleanly, rather than the pollution-prone 0.15 to 0.25 equivalence ratio range which would occupy a homogenous IPC engine’s combustion chamber. Note that the volume of the perimeter squish region approaches zero at TDC, whereas the volume of the central combustion region approaches a finite value at TDC, creating an effective air pump directed from the perimeter squish region toward the central combustion region in the last 12mm before TDC.

The central combustion region isshaped to fully support combustion: The surface area of the central combustion region is comparatively low to assist a speedy combustion reaction. The insulated chamber surface heats up quickly during compression and combustion to assure fuel in close proximity to the insulated material combusts properly. The central combustion region isshaped to generate within itself a toroidal vortex as air is pumped in from the perimeter squish region, assuring all fuel is in motion to uniformly combust, the turbulence minimizing both cold and hot spots in the central combustion region which helps prevent pre-ignition.

The annular transfer passage acts to buffer combustion at TDC. As the combusting reaction heats up at TDC, it expands beyond the central combustion region. The combusting gasses efficiently spill into a segment of the annular transfer passage which fully supports combustion, while pure air residing within the annular transfer passage is pushed into the perimeter squish region. Only when the piston falls 0.5mm after TDC do combusted gassessignificantly occupy the annular transfer passage and begin to approach the perimeter squish region. By this time the combustion reaction has been consumed and concluded.

Any residual fuel that is not completely combusted when the piston falls to 0.5mm ATC will exit the combustion chamber as a pollutant. There is not a second opportunity to combust fuel that does not combust near TDC. If exhaust emissions are to be low, creviced features, such as valve seats and spark plug insulation recesses, are not permissible in the combustion area. The central combustion region at TDC is sized to be half the volume of the perimeter squish region plus the backfill passage at TDC, allowing a full throttle fuel-air equivalence ratio of 0.80.


Compression Ratio and Expansion Ratio
The IPC engine inducts unthrottled air, much like a Diesel engine. The IPC engine adiabatically pre-warms the induction charge during compression to just below the auto-ignition temperature of the fuel-air mixture, promoting rapid combustion when a spark is generated near TDC. This puts the dynamic compression ratio (DCR) at roughly 15:1. In flex-fuel configurations, the compression ratio is actively regulated to assure compression pressure remains just below the autoignition level as conditions change. This is accomplished by monitoring ignition reactivity and continuously servoing valve closure timing to suit.

The dynamic expansion ratio (DER) will be about 30:1 to minimize heat energy loss to the exhaust duct, much the way an Atkinson engine minimizes exhaust energy loss. The selection of 30:1 for the DER is based on the assumption that a peak combustion chamber pressure of 150 bar at TDC will not form oxides of nitrogen pollutants, and on the prevalence of predominantly diatomic gasses of the fuel-lean combusted charge obeying, to a first order approximation, the 150 bar / (30 ^ 1.4) = 1.3 bar equation. Mechanical friction drives a deviation from the 1.3 bar specification at BDC, though fuel prices may additionally influence the selected expansion ratio. An unconventionally large expansion ratio is chosen to extract virtually all useable heat and pressure from the combustion chamber before the exhaust valve opens, resulting in a comparatively cool and quiet exhaust cycle with minimal exhaust duct flow velocity.

The DCR can be referred to as the "compression ratio", and the DER referred to as the "expansion ratio". If the 4-stroke IPC engine has a 100mm piston stroke, the expansion cycle occupies 100mm of piston travel after TDC, and the 15:1 compression cycle begins 50mm BTC. The 15:1 compression ratio is independent of the 30:1 expansion ratio, and each can be adjusted as required.

The induction cycle for a deluxe version of the 4-stroke IPC engine occupies only the first 50mm of piston travel after TDC and the compression cycle occupies the final 50mm of piston travel before TDC. Combustion chamber pressure will drop as low as 0.50 ^ 1.4 = 0.38 bar in the period between the end of the induction cycle and the start of the compression cycle. The described induction cycle requires a valve train configuration with an unusually large camshaft base circle, in order to actuate the valves through such a small camshaft arc. An intermediate version of the 4-stroke IPC engine reduces cost and complexity by incorporating the Atkinson reversion cycle, in which the induction cycle occupies the entire 100mm of piston travel from TDC to BDC, and as the piston then rises from BDC the inducted air flows backward out the intake duct until the intake valve closes at 50mm BTC.



Fig25 – Perimeter squish region (formerly called the crevice chamber), central combustion region (formerly called the central combustion chamber), and annular transfer passage (formerly called either the annular passage or alternately the backfill passage), are formed at 12mm BTC.


4-Stroke IPC Engine Sequence

The deluxe 4-stroke IPC engine cycle includes twelve stages of operation including: 1) Intake, 2) Vacuum, 3) Rebound, 4) Compression, 5) Injection, 6) Turbulence, 7) Ignition, 8) Combustion, 9) Expansion, 10) Vacuum, 11) Rebound, and 12) Exhaust. The engine cycle includes the following sequence:
04mm ATC: Intake valve opens, drawing in unthrottled air, same as a Diesel engine.
50mm ATC: Induction cycle ends, intake valve closes.
51mm ATC: Cylinder begins pulling a vacuum as piston continues toward BDC.
100mm BDC: Combustion chamber drops to 0.50 ^ 1.4 = 0.38 bar pressure.
99mm BTC: Piston elastically rebounds off vacuum and is pulled toward TDC.
50mm BTC: Vacuum rebound ends, compression of inducted air begins.
49mm BTC: Inducted air begins adiabatically heating in combustion chamber.
12mm BTC: Combustion chamber transitions to become stratified.
09mm BTC: Combustion chamber becomes predominantly thermally insulating.
08mm BTC: Fuel is direct injected toward pocket at center of piston.
07mm BTC: Perimeter squish region pumps fresh air toward piston pocket, constraining fuel.
06mm BTC: Direct fuel injection ends.
05mm BTC: Air from perimeter squish region generates turbulence in central combustion region.
01mm BTC: Fuel and air homogenously mixed in central combustion region.
0.5mm BTC: Spark ignites fuel and air mixture, combustion progresses rapidly.
00mm TDC: Combustion reaction expands into annular transfer passage.
0.3mm ATC: Annular transfer passage forces pure air back into perimeter squish region.
0.5mm ATC: Combustion reaction extinguishes.
05mm ATC: Combusted gasses are adiabatically cooling in combustion chamber.
09mm ATC: Combustion chamber first exposes thermally conductive cylinder bore.
12mm ATC: Stratified combustion chamber transitions to become single chamber.
75mm ATC: Combustion chamber starts pulling a vacuum (low throttle only).
87mm ATC: Combustion chamber starts pulling a vacuum (mid throttle only).
99mm ATC: Combustion chamber pressure drops to 1.3 bar (full throttle only).
100mm BDC: Combustion chamber pressure or vacuum depends on throttle position.
99mm BTC: Expansion cycle ends, exhaust valve opens (full throttle only).
87mm BTC: Combustion chamber vacuum ends, exhaust valve opens (mid throttle only).
75mm BTC: Combustion chamber vacuum ends, exhaust valve opens (low throttle only).
04mm BTC: Exhaust valve closes.
04mm ATC: Intake valve opens, drawing in unthrottled air, same as a Diesel engine.

The deluxe 4-stroke IPC engine described above has a rather complex valve train. The deluxe 4-stroke IPC engine can be cost-reduced to employ a simpler, slightly less thermally efficient Atkinson reversion cycle. This intermediate version of the 4-stroke IPC engine follows the sequence:
04mm ATC: Intake valve opens, drawing in unthrottled air, same as Diesel engine.
100mm BDC: Induction cycle ends, Atkinson reversion cycle begins.
50mm BTC: Intake valve closes, Atkinson reversion ends, compression cycle begins.
49mm BTC: Fresh air begins adiabatically heating in combustion chamber.
12mm BTC: Combustion chamber transitions to become stratified.
09mm BTC: Combustion chamber becomes predominantly thermally insulating.
08mm BTC: Fuel is direct injected toward pocket at center of piston.
07mm BTC: Perimeter squish region pumps fresh air toward piston pocket, constraining fuel.
06mm BTC: Direct fuel injection ends.
05mm BTC: Air from perimeter squish region generates turbulence in central combustion region.
01mm BTC: Fuel and air homogenously mixed in central combustion region.
0.5mm BTC: Spark ignites fuel and air mixture, combustion progresses rapidly.
00mm TDC: Combustion reaction expands into annular transfer passage.
0.3mm ATC: Annular transfer passage forces pure air back into perimeter squish region.
0.5mm ATC: Combustion reaction extinguishes.
05mm ATC: Combusted gasses are adiabatically cooling in combustion chamber.
09mm ATC: Combustion chamber first exposes thermally conductive cylinder bore.
12mm ATC: Stratified combustion chamber transitions to become single chamber.
50mm ATC: Conventional expansion cycle ends, Atkinson expansion cycle begins.
100mm BDC: Atkinson expansion cycle ends, exhaust valve opens, exhaust cycle begins.
04mm BTC: Exhaust valve closes, exhaust cycle ends.
04mm ATC: Intake valve opens, drawing in unthrottled air, same as Diesel engine.



2-Stroke IPC Engine Sequence

A deluxe 2-stroke IPC engine incorporates an engine operating sequence summarized as follows:
1) Compression - 33mm BTC to 0.5mm BTC
2) Ignition – 0.5mm BTC
3) Combustion – 0.5mm BTC to 0.5mm ATC
4) Expansion – 0.5mm ATC to 67mm ATC
5) Induction - 67mm ATC to 90mm BTC
6) Exhaustion - 90mm BTC to 33mm BTC


A deluxe 2-stroke IPC engine includes intake ports on the cylinder bore and exhaust poppet valves in the head, and the operating sequence includes:
33mm BTC: Exhaust valve closes, compression of fresh air and some exhaust begins.
32mm BTC: Fresh air begins adiabatically heating.
12mm BTC: Combustion chamber transitions to become stratified.
09mm BTC: Combustion chamber becomes predominantly thermally insulating.
08mm BTC: Fuel is direct injected toward pocket at center of piston.
07mm BTC: Perimeter squish region pumps fresh air toward piston pocket, constraining fuel.
06mm BTC: Direct fuel injection ends.
05mm BTC: Air from perimeter squish region generates turbulence in central combustion region.
01mm BTC: Fuel and air homogenously mixed in central combustion region.
0.5mm BTC: Spark ignites fuel and air mixture, combustion progresses rapidly.
00mm TDC: Combustion reaction expands into annular transfer passage.
0.3mm ATC: Annular transfer passage forces pure air back into perimeter squish region.
0.5mm ATC: Combustion reaction extinguishes.
05mm ATC: Combusted gasses are adiabatically cooling in combustion chamber.
09mm ATC: Combustion chamber first exposes thermally conductive cylinder bore.
12mm ATC: Stratified combustion chamber transitions to become single chamber.
33mm ATC: Conventional expansion cycle ends, Atkinson expansion cycle begins.
66mm ATC: Combustion chamber pressure reaches 1BAR.
67mm ATC: Intake port becomes visible to combustion chamber.
68mm ATC: Vacuum forms and pulls fresh air into lower third of combustion chamber.
69mm ATC: Upper 67mm of chamber contains gasses with ¼ of oxygen consumed.
90mm ATC: Exhaust valves in head begin to open.
100mm BDC: Intake ports are fully visible to combustion chamber.
99mm BTC: Lower 33mm of combustion chamber contains air, upper 67mm contains exhaust.
90mm BTC: Intake ports in cylinder bore become blocked by rotating drum valve assy.
89mm BTC: Piston pushes combusted gasses in upper chamber into exhaust duct.
33mm BTC: Exhaust valves close, compression of fresh air and some exhaust begins.

An intermediate version of the 2-stroke IPC engine employs rotary drum valves for both induction and exhaustion, with intake and exhaust ports located on the cylinder bore, and the engine operates in a sequence summarized as follows:
1) Compression - 33mm BTC to 0.5mm BTC
2) Ignition – 0.5mm BTC
3) Combustion – 0.5mm BTC to 0.5mm ATC
4) Expansion – 0.5mm ATC to 67mm ATC
5) Simultaneous Induction and Exhaustion - 67mm ATC to 33mm BTC


This intermediate version of the 2-stroke IPC engine contains no poppet valves in the head. Instead, this intermediate 2-stroke IPC engine uses intake ports on one side of the cylinder bore, and exhaust ports on the opposite side of the cylinder bore. The intake ports are organized into an upper bank of ports and a lower bank of ports in which the rotary drum valve acts as a shutter and also acts as a blower. The exhaust side of the cylinder bore issimilarly configured, except the rotary drum valve assembly acts as a vacuum pump. Induction and exhaustion occur simultaneously, flowing across the combustion chamber with sufficient chaos that combusted gasses throughout the combustion chamber are substantially replaced with inducted air. The detailed operating sequence is as follows:
33mm BTC: Exhaust port sealed by piston ring, compression begins.
32mm BTC: Fresh air begins adiabatically heating.
12mm BTC: Combustion chamber transitions to become stratified.
09mm BTC: Combustion chamber becomes predominantly thermally insulating.
08mm BTC: Fuel is direct injected toward pocket at center of piston.
07mm BTC: Perimeter squish region pumps fresh air toward piston pocket, constraining fuel.
06mm BTC: Direct fuel injection ends.
05mm BTC: Air from perimeter squish region generates turbulence in central combustion region.
01mm BTC: Fuel and air homogenously mixed in central combustion region.
0.5mm BTC: Spark ignites fuel and air mixture, combustion progresses rapidly.
00mm TDC: Combustion reaction expands into annular transfer passage.
0.3mm ATC: Annular transfer passage forces pure air back into perimeter squish region.
0.5mm ATC: Combustion reaction extinguishes.
05mm ATC: Combusted gasses are adiabatically cooling in combustion chamber.
09mm ATC: Combustion chamber first exposes thermally conductive cylinder bore.
12mm ATC: Stratified combustion chamber transitions to become single chamber.
33mm ATC: Upper intake and exhaust ports first enter chamber but are shuttered closed.
50mm ATC: Upper intake and exhaust ports remain shuttered but are fully in chamber.
67mm ATC: Lower intake and exhaust ports first become exposed to chamber.
68mm ATC: Cross-flow of inducted and exhausted gasses begins in chamber.
69mm ATC: Upper ports begin to unshutter and begin to cross-flow.
90mm ATC: All intake and exhaust ports are now unshuttered and cross-flowing.
100mm BDC: Lower intake and exhaust ports fully visible to combustion chamber.
90mm BTC: Intake ports become blocked by rotary drum valve assembly.
89mm BDC: Piston pushes combustion chamber gasses out exhaust ports.
33mm BTC: Exhaust port sealed by piston rings, compression begins.

A simplified version of the 2-stroke IPC engine takes the previously described "opposing" intake and exhaust ports and places them on the same side of the cylinder block, with a single rotary drum valve assembly modified to provide both induction and exhaustion, and with the externally visible surfaces of the rotary drum valve assembly acting on the induction gasses and the internally visible surfaces of the rotary drum valve assembly acting on the exhaust gasses via a window in the drum. This results in a low-cost 2-stroke IPC engine.



Fig26 – Annular transfer passage (formerly called the backfill passage) has three times the volume of the perimeter squish region (formerly called the crevice chamber) at TDC.



Fig27 – 8 valves per cylinder is the simplest theoretical arrangement for the 4-stroke IPC engine. It should be remembered this is a theoretically ideal arrangement with application during early prototyping. Advanced computational analysis will cost-reduce this construction once the ideal arrangement is found to be valid. The intermediate and simplified versions of the 2-stroke IPC engine require no valves or camshafts, making for simple, low cost prototypes which may carry the economy of this combustion chamber directly into production.





Friction reduction
Mechanical friction and windage friction are significant issues in a low volumetric efficiency engine like the IPC engine. Presented here are some of the friction reducing methods being considered for this engine:

Twin gear-synchronized counter-rotating crankshafts can eliminate piston side thrust forces, eliminating skirt thrust friction and piston slap. Twin crankshafts are presently implemented in the IPC engine in order to learn about their characteristics, but the next stage of development will eliminate them due to cost, weight, and size constraints.



If twin counter-rotating crankshafts are chosen, resulting in the elimination of piston side thrust forces, the piston skirt can be sculpted to reduce contact surface area with the cylinder bore without concern for skirt wear, resulting in reduced piston skirt viscous oil friction losses.

In single crankshaft configurations, long connecting rods and anti-friction material applied to piston skirt will each reduce piston side thrust friction. When considering the short duration of high cylinder pressure after TDC, skirt friction need not be a significant issue in a single crankshaft version of the IPC engine.


Since the insulated pulse engine has high cylinder pressures for a very short percentage of crankshaft rotation, gas ported low-tension piston rings reduce piston sliding friction through the bulk of the piston travel while sealing tightly when pressure is high. The IPC engine’s combustion chamber is transitionally stratified and permits only pure filtered air to enter the piston’s gas ports, preventing localized HC exhaust emissions and also preventing gas port clogging. Presently, there are no production engines which can use gas ported pistons and achieve low exhaust emissions, limiting application only to racing engines. Since friction minimization is a primary consideration in the success of any low volumetric efficiency engine, the IPC engine may find gas ported pistons crucial to achieving commercial applicability.


As noted in the previous two paragraphs, the combination of short-duty skirt thrust forces and short-duty for high ring tension, may provide a unique opportunity to significantly reduce friction in the IPC engine.

Intake port friction reduction: Intake air in the IPC engine is unthrottled, as it is in Diesel engines. Throttled intake air, as found in Otto engines, consumes more pumping energy than unthrottled intake air.
Exhaust port friction reduction: There is minimal opportunity for “tuning” the exhaust with headers, since there isso little energy in the expelled exhaust. The exhaust ports are simply made short and free-flowing, and dump into an exhaust plenum that directs to a low restriction outlet pipe. Since the exhaust is comparatively cool, and since the exhaust flow velocity is low and contains minimal kinetic energy, it is expected the IPC engine will need little or no muffling.












Fig32 – Bulkhead between cyls 1&2 and 3&4 are vented to reduce crankcase turbulence in this 4-stroke IPC engine. This engine runs smoothly without counterweights or countershafts, allowing the use of unusually lightweight nodular iron crankshafts.


Combustion chamber turbulence friction reduction: Turbulence is preferably generated only from time direct fuel injection commences until the instant combustion extinguishes, to optimally pre-mix fuel and air and promote rapid combustion. Turbulence generated significantly prior to fuel injection, such as conventional tumble or swirl configurations found in many engines today, can be considered wasted vortex energy. Turbulence generated after the combustion process extinguishes is also wasted vortex energy. The IPC engine’s construction targets this goal as close as possible.

Crankcase turbulence friction reduction: Block bulkhead ports between pistons 1&2 and pistons 3&4 relieve an air pressure transient in crankcase caused when neighboring pistons travel in opposite directions. A crankcase vacuum pump may be employed to reduce crankcase windage friction.



Unfinished topics to be included in this draft



With exception of the intermediate and simplified versions of the 2-stroke, the IPC engine is an “interference engine” by design. The shown timing belt configuration should be replaced in the interference-type engines by a timing chain, since timing chains typically need no replacement during the service life of an engine.


An overpressure bypass valve, as seen in early IPC combustion chamber images, has been omitted in favor of an IPC combustion chamber fortified to withstand occasional stoichiometric combustion events, much the way ordinary Otto engine combustion chambers are fortified to sustain occasional detonation events.


A starter motor and alternator are not part of the presented IPC engine construction, as the target application is a flywheel-electric hybrid motor vehicle which multitasks the traction drive electric motor for traction drive, engine starting, and also conventional battery charging. This application uses an electrically-interfaced 500kJ/25kg carbon flywheel module for short-term traction energy storage, with flywheel energy efficiently managed through GPS terrain maps and driving history.




Limited Description of the 2-Stroke Insulated Pulse Engine

A conceptual 3.2 liter in-line 4-cylinder 2-stroke IPC engine is presented here in limited detail, since the model is now obsolete and being updated. Asample image is located at the top of this web page, and more image samples are presented below. The 3.2 liter 2-stroke is expected to produce roughly 50 horsepower at 4000 RPM, and exhaust emissions are expected to be comparable to the 4-stroke IPC engine, combusting cleanly with little or no need for emissions controls.




Introductory Image of 2-Stroke Shortblock Shown below in Fig 37 are eight rotary drum valves (green) which act as both a maintenance-free reed valve and also as a silent replacement for a roots type blower. The shutter drum valve has a uniform clearance of 0.5mm between the rotor and the block and cylinder to eliminate friction and wear, as there is no need for perfect sealing. The shutter drum valve, when shuttered closed across the intake port as the piston rises, need only apply significantly more resistance to flow than the opened exhaust valves generate, not perfect resistance to flow. As shown, each of the two shutter drum valve assembly prototypes contain six sealed ball bearings. This will be cost-reduced to three bearings per assembly when eventually stress modeled.

Prior to the shutter drum valve closing off the intake ports, the shutter drum valve scoops air ahead of it, providing a calibrated volume of positive-pressure fresh air which will assist filling the lower 1/3 of the combustion chamber as the piston falls toward BDC. The combustion chamber begins pulling a vacuum when the piston falls to 67mm ATC, as that is when the combustion reaction drops to 1 ATM pressure. When the piston reaches BDC, the top 2/3 of the combustion chamber contains exhaust, and the bottom 1/3 contains fresh air. As the piston rises, it pushes the top 2/3 combusted gas out the exhaust valves, sealing in the bottom 1/3 of fresh air for the next combustion reaction.

The shutter drum valve at the intake port differs from a roots type blower in that the shutter drum valve works to displace gasses only during the portion of the engine cycle in which the cylinder ports are exposed to the combustion chamber and able to flow, thus eliminating roots-type blower losses related to pumping against plenum pressure. Each shutter drum valve may be accompanied by a supplemental rotor, much as a roots blower uses two rotors to displace gasses, if more control of transfer volume is sought. Due to the lack of head pressure, the supplemental rotor need not contact the shutter drum valve, and therefore will be maintenance free, but should be constructed to assure the small clearance between rotors operates at tolerable leakage levels. In the intermediate version of the 2-stroke IPC engine an additional shutter drum valve applied to the exhaust port may be used to draw combusted gasses from the combustion chamber. The shutter drum valve differs from a reed valve in that it does not cyclically flex during operation and therefore requires no maintenance during the lifetime of the engine.

Since the IPC engine at full throttle consumes only 1/4 of the compressed oxygen each combustion event, the exhaust retains 75% unconsumed oxygen. Partial mixing of exhaust with inducted air while the piston rises from BDC will have minimal effect on the combustion reaction.


Fig 37 - Introductory image of twin-crankshaft shortblock for 2-stroke IPC engine. The study of twin crankshafts has concluded, and single crankshaft studies are now proceeding.


Fig 38 - Numerous aging images have been removed from this paper to improve clarity. This paper will present a full set of updated images when they become available.



End of draft of 18Apr2010, updated 04Dec2010


The Insulated Pulse Engine: A cold adiabatic engine concept
By Dave Schouweiler, shoe@bitstream.net, Minneapolis MN USA, updated 04Dec2010


2

A cold adiabatic engine concept by Dave Schouweiler, shoe@bitstream.net, insulatedpulseengine.com.
This project describes a thermally efficient concept for combusting fuel in an internal combustion engine. It explores adiabatic engines, a usually dormant science that was last active for several years after the 1979 oil crisis. This "cold adiabatic engine” concept is not like published adiabatic engines which expel superheated combustion gasses into an exhaust duct for the post-processing of energy. Instead, this engine allows combusted gasses to adiabatically expand and cool before exiting the combustion chamber.
This research is a work in progress. This draft is a conceptual paper, not a technical paper, as it contains intuitive approximations and primitive constructions which require refinement. The images exist as props to showcase the concept, not to suggest a best design practice. I’m not an engine designer, but seek to learn why this concept may or may not work. Technical critique and information on similar experiments is welcome.
Brief Introduction to the Insulated Pulse Engine

The "insulated pulse combustion engine", abbreviated "insulated pulse engine" or "IPC engine", is a low volumetric efficiency internal combustion engine concept which combusts fuel at high thermal efficiency. The combustion chamber is selectively insulated to minimize heat energy loss to a cooling system. Combustion initiates and is consumed rapidly near top dead center (TDC), permitting adiabatic cooling of combustion chamber gasses through the entire expansion stroke. The expansion stroke is also extended, further reducing average combustion chamber temperatures, to minimize stress on the thermal insulators and to extract additional heat and pressure from the combustion chamber, resulting in an exhaust that is comparatively cool and pressureless.
When compared with Otto and Diesel engines at full throttle, a similarly displaced IPC engine at full throttle consumes roughly an eighth of the fuel each combustion cycle. As will be described in this paper, the IPC engine is expected to have twice the fuel efficiency of Otto and Diesel engines, and will therefore generate a fourth of the horsepower of similarly displaced Otto and Diesel engines at full throttle and similar RPM. This paper constructs a hypothetical 3.2 liter in-line 4-stroke IPC engine that generates an expected 50 HP at 4000 RPM (Fig10). This paper ends with discussion of a hypothetical 3.2 liter in-line 2-stroke IPC engine (images are now being created).
The cylinder displacement of an IPC engine is roughly four times that of Otto and Diesel engines at equivalent horsepower and RPM, but the cost, weight, and volume of the IPC engine assembly remains comparable due to a reduction in need for cooling, muffling, and emissions control components. Since mechanical friction is a variable which correlates more closely to generated horsepower than to displacement, and since the IPC engine is constructed using methods which emphasize reduction of mechanical friction and windage friction, friction generated within the IPC engine is comparable to friction generated within equivalently powered Otto and Diesel engines.
Fig10 – IPC engine, PTO flange at lower left, cyl#4 head cutaway shown upper left.




Volumetric Efficiency and Thermal Efficiency

The term “volumetric efficiency” is often used to describe the horsepower attributes of an engine. The term “thermal efficiency” is often used to describe the fuel consumption attributes of an engine. Lots of progress has been made with volumetric efficiency in engines over the decades. Far less progress has been made with thermal efficiency. When “improved efficiency” is mentioned in relation to an internal combustion engine, it is usually “volumetric efficiency” that is being discussed, even when referencing improved fuel mileage. This can cause confusion between the terms.
Automobile manufacturers are often able to improve fuel mileage by better matching an engine to a vehicle. It was formerly practical to have a large displacement engine in a vehicle with plenty of reserve power on tap, but large displacement engines have more combustion chamber surface area exposed to hot combustion chamber gasses than do small displacement engines, and this supports the loss of combustion heat energy to the cooling system, increasing costs to propel the vehicle.
Engines for a particular application now tend to be of smaller displacement and higher volumetric efficiency, allowing them to generate the horsepower levels of older large displacement engines. Higher volumetric efficiency in modern engines does not necessarily indicate improved thermal efficiency in an engine. For instance, an older 80 horsepower 2 liter engine and a modern 160 horsepower 2 liter engine likely provide about the same fuel mileage in a particular small car. Higher volumetric efficiency does enable a more thermally efficient matching of small engine to large vehicle. A small displacement engine must generally operate at higher cylinder pressures to make the same power as a large displacement engine, and higher cylinder pressures in an engine are often more thermally efficient than lower cylinder pressures in the same engine.
Cooling System Losses

Internal combustion engines incorporate a cooling system to quickly remove heat absorbed by combustion chamber metals after each combustion event. This removal is necessary, since chamber metals would otherwise attain the average temperature of the combustion chamber gasses, a temperature too hot in Otto and Diesel engines for sustainable engine operation.
Following the oil crisis of 1979, engine manufacturers around the world began developing internal combustion engine prototypes which contained thermally insulated combustion chambers in an attempt to improve engine thermal efficiency without sacrificing volumetric efficiency. Thermally insulating the combustion chamber with ceramic inserts reduced, and sometimes eliminated, the need for a cooling system, and thus retained a larger fraction of combustion heat energy for mechanical work output. Experimental results on three of these ceramic engine or "adiabatic engine" projects can be reviewed in SAE technical papers 810070 (1981), 820431 (1982), and 840428 (1984), whose abstracts are currently viewable at www.sae.org/technical/papers, and where the papers may be downloaded.
Stressful operating conditions limited the durability of the brittle ceramic insulators. Adiabatic engine prototypes of the 1980s did not find commercial success. The use of ceramics to insulate combustion chambers of internal combustion engines for the primary purpose of improving fuel mileage in vehicles has found minimal research interest in the industry since the conclusion of these experiments.
Fig11 – Low friction internals: Twin crankshafts allow reduced piston skirt contact area.




Basic Description of the Insulated Pulse Engine

The insulated pulse engine is an ordinary reciprocating piston, spark-ignition, engine which applies three basic yet unconventional features to achieve high thermal efficiency:
  1. Insulated combustion chamber (like adiabatic engine)
  2. Rapid "pulse" combustion (like HCCI engine)
  3. Extended expansion cycle (like Atkinson engine)
These three features create an engine with both high thermal efficiency and low volumetric efficiency. Mechanical friction and windage friction become significant issues in a low volumetric efficiency engine, so the insulated pulse engine must also apply ordinary but unconventional methods to reduce friction:
  1. Twin counter-rotating crankshafts eliminate piston side thrust friction (Figs11, 13)
  2. Reduced piston skirt contact area reduces piston skirt oil viscosity friction
  3. Gas ported low-tension piston rings reduce piston sliding friction (Fig12)
  4. Minimize port flow volume, port flow resistance, and port flow turbulence
  5. Minimize crankcase windage (air turbulence and air entrapment, Figs31, 32)
  6. Assure combustion chamber turbulence efficiently mixes fuel with air and is present during ignition and combustion.
The resulting engine requires only an active oil cooler of ordinary capacity to support all cooling needs and does not require a muffler to function quietly.
Fig12 – Gas ported piston with low tension compression rings and reduced skirt contact.


Fig13 – Conceptual piston and rod assembly for dual crankshaft construction.




In both Otto and Diesel engines, combustion is engineered to progress gradually, beginning near TDC and continuing well into the expansion cycle. This low heat release rate allows a lot of fuel to burn without reaching the pressure limits of the combustion chamber, providing high volumetric efficiency and low thermal efficiency. Thermal efficiency is low because the late burning fuel cannot adiabatically expand as many times as the early burning fuel. This causes large amounts of energy to be lost to the exhaust in the form of heat and pressure.
Adiabatic engines of the 1980s similarly combusted with a low heat release rate, resulting in high volumetric efficiency and low thermal efficiency, but differed in that they incorporated ceramic insulators to superheat the combustion chamber by preventing heat escape to a cooling system. The superheated combustion chamber then superheated the exhaust duct for the purpose of energy recovery through turbocompounding and other heat recovery methods. Adiabatic engines of the 1980s operated under the most brutal conditions. They could not be made sufficiently reliable for commercial application.
In both HCCI and IPC engines, combustion initiates near TDC and is rapidly consumed near TDC, providing combustion with low volumetric efficiency and high thermal efficiency. The volumetric efficiency is low because a comparatively small amount of fuel will generate sufficient temperature and pressure near TDC to reach the limits which do not form oxides of nitrogen exhaust pollutants. Thermal efficiency is high because all of the combusted gasses adiabatically cool through the entire expansion stroke, lowering the average temperature of the combustion chamber and greatly reducing the percentage of heat energy loss to the exhaust. The ordinary methods selected to achieve this high heat release rate are:
  1. Fuel-air charge fully mixed prior to ignition
  2. High compression ratio
  3. Combustion chamber turbulence during ignition and combustion
  4. Fuel-lean stoichiometry
The HCCI engine uses compression ignition and a homogenous fuel-lean charge to ignite a dozen or so degrees before TDC. The IPC engine must use spark ignition and a stratified fuel-lean charge to ignite just a few degrees before TDC. The IPC engine must actively regulate the compression ratio to maximize compression pressure while preventing autoignition, and it does so by monitoring ignition reactivity and continuously servoing intake valve closure timing. The later ignition timing of the IPC engine is required to provide sufficient turbulence for complete fuel/air mixing, but the later ignition also reduces bearing load durations, thus momentarily reducing friction, and it further reduces the average combustion chamber surface temperature to a level which is less stressful on combustion chamber materials while preserving heat energy for mechanical work output.
Fig14 – Piston shown at 030BTC, just as combustion chamber becomes fully insulating.




The IPC engine thermally insulates the combustion chamber completely when the piston is within 30 degrees of TDC (Fig14), and partly insulates when the piston is further than 30 degrees from TDC. The goal is to minimize combustion chamber heat energy loss to the cooling system during the hottest portion of the compression and expansion cycles. The combustion chamber must additionally be designed to generate turbulence which adiabatically pre-heats critical chamber surfaces while locally mixing the direct injected fuel with air in the stratified central chamber just prior to ignition to assure complete combustion.
The preferred thermal insulating material in the IPC engine is a steel alloy with 40% nickel and 3mm thickness with thermal conductivity of 10 W/m K at 200 degrees C. A PSZ ceramic of 3mm thickness and thermal conductivity of 2 W/m K is reserved as a potential insulator in a future state of development of this engine concept, but more research is needed before ceramic can be proven reliable. As a comparison, the thermal conductivity of cast A356-T6 aluminum is 130 W/m K at 200 degrees C with typical thermal gradient distance of 10mm between combustion chamber and cooling system, and compacted gray iron is 40 W/m K with typical gradient distance of 5mm.
Exhaust emission concerns in the IPC engine fall into four simplified categories:
  1. Hydrocarbon exhaust emissions, representing fuel that is not combusted, are formed when fuel is in proximity of chilled combustion chamber crevices such as are found near the head gasket, upper piston ring, and intake valve seat.
  2. Soot emissions, representing fuel that is 1/3 combusted, are formed when fuel is direct injected into the dense flame kernel of a compression ignition engine which has already consumed all adjacent oxygen.
  3. Carbon monoxide emissions, representing fuel that is 2/3 combusted, are formed when fuel is combusted near chilled surfaces within the combustion chamber.
  4. Oxides of nitrogen emissions are generated when heat energy becomes unnecessarily high in the combustion chamber and the very stable 3-bond nitrogen molecule breaks apart.
It is recognized there is nothing simple about exhaust emissions in any internal combustion engine. Exhaust emissions cannot be more deeply addressed in the IPC engine concept until the most basic thermochemical and energy equations are completed. This work will commence when the conceptual mechanical CAD work is completed and presented in this paper. Once the energy calculations are done, there may be opportunity to computationally simulate the IPC engine using specially developed software which will step closer and closer to a functioning engine, eventually offering potential to theoretically prove or disprove this concept on paper. To be sure, there is a lot to study, calculate, and learn from each stage of this evolving process.
The unusually shaped insulated components of the IPC combustion chamber form a significant piston crevice volume above the compression sealing rings between the piston and cylinder bore. This calls for incorporation of a transitional annular chamber, called a crevice chamber, in conjunction with a unique direct fuel injection algorithm, to actively prevent fuel from entering the chilly crevice gaps, thus preventing crevice-induced hydrocarbon emissions. This crevice chamber will act as an annular air pump which rejects the admission of injected fuel into peripheral areas of the combustion chamber which contain crevice gaps. As an annular air pump, the crevice chamber generates flow of adiabatically heated air across critical combustion chamber surfaces during compression, prewarming the surfaces to reduce carbon monoxide emissions. The the air flow continues into the central combustion chamber where it generates toroidal turbulence which efficiently premix fuel with air, thus eliminating soot emissions. It is expected the IPC engine will generate very few exhaust pollutants and will require no emissions control devices.
The IPC incorporates an extended expansion cycle, much like an Atkinson engine, to reduce pressure and heat energy levels in the exhaust duct while further reducing the average combustion chamber temperature, thus bringing the average combustion chamber temperature down to a point where a cooling system is not required at all, except perhaps when running in hot ambient conditions. When cooling is required, the requirements are sufficiently slight that an external oil cooling system of ordinary capacity is capable of dissipating the absorbed combustion energy.
The extended expansion cycle further reduces volumetric efficiency while increasing thermal efficiency, but, as mentioned in the introduction, the IPC engine makes up for volumetric efficiency reductions by substantially increasing cylinder displacement. The extended expansion cycle singularly causes displacement of the IPC engine to roughly double to retain comparable power output. An expansion ratio value is selected which assures expansion energy gains constructively exceed friction force losses through the entire expansion stroke.
The IPC engine expels a comparatively cool exhaust. This differentiates it from published adiabatic engines, which usually run without a cooling system at the expense of superheating the combustion chamber and exhaust duct. These published adiabatic engines always incorporated turbocharging, turbocompounding, or other post-processing methods to supplement fuel energy conversion efficiency.
Insulation Efficiency, Combustion Efficiency, and Friction Efficiency
Thermal efficiency in an internal combustion engine is recognized in three basic forms:
  1. Insulation efficiency reduces the conducted loss of combustion energy to a cooling system in the form of heat. High insulation efficiency is one of two basic elements found a true adiabatic engine.
  2. Combustion efficiency reduces the convected loss of combustion energy to the exhaust duct in the form of heat and pressure. High combustion efficiency is the second of two basic elements found in a true adiabatic engine.
  3. Friction efficiency reduces combustion energy loss to mechanical friction and air pumping.
Insulation efficiency was incorporated into the adiabatic engine experiments of the early 1980s, but combustion efficiency was not. These “adiabatic engines” were, in effect, half-adiabatic, not fully adiabatic. These experiments retained a low heat release rate which generated significant heat energy loss to the exhaust cycle. Only the fuel burning near TDC combusted at high adiabatic efficiency. The bulk of the fuel was combusted after TDC had passed, and it combusted at reduced adiabatic efficiency. The result was a brutally hot combustion and exhaust process which provided some improvement in thermal efficiency over Otto and Diesel engines, but did not allow sufficient reliability for commercial applicability.
Combustion efficiency is incorporated into the HCCI prototype engines being researched around the world today, as the entire combustion reaction occurs near TDC, but insulation efficiency is not. This defines the HCCI engine as a different type of “half-adiabatic” engine, since it transfers heat energy to a cooling system and a fully adiabatic engine does not.
Insulation efficiency and combustion efficiency are both incorporated into the IPC engine, and the constructions described below will provide a notable increase in fuel efficiency over both adiabatic and HCCI engines. The IPC engine is a true adiabatic engine construction, but to prevent confusion with established naming practice, the IPC engine is probably best called a “cold adiabatic engine”, since it transmits minimal heat to a cooling system and expels minimal heat energy into the exhaust duct.
Detailed Description of the Insulated Pulse Engine

The most significant feature of the IPC engine is the thermally insulated combustion chamber: The piston contains an insulating cap (Figs15, 16), and the cylinder head contains an insulating dish (Figs 16, 17). This is the full extent of the insulation, cheap and simple. On a preliminary note, the unique size and shape of the combustion chamber has forced a reduction of the valve head diameter, requiring a multi-valve arrangement in order to retain a low-restriction intake and exhaust flow. This multi-valve configuration will be discussed later in this paper. These two insulating components will be investment cast out of a tough, cheap, nickel steel alloy chosen for its low thermal conductivity, low thermal expansion coefficient, high temperature stability, and valve seat wear resistance. For valve seat wear resistance, there may be some carbon added to permit localized induction hardening. One of these insulators will be insert-molded into a die cast aluminum piston to keep reciprocating mass low, the other will be insert molded into a cast aluminum cylinder head to keep engine mass low. The head insulator will economically combine the duties of valve seat, spark/injector mount, overpressure protection valve mount, and head gasket sealing surface.
Fig15 - Insulating piston cap, made from investment cast 40% nickel steel.


Fig16 – Back view of piston cap and head dish showing insert-casting retention features.


Fig17 – Insulating 40% nickel steel head dish with thermal conductivity of 10 W/m K.




The valves installed into this head will be made of an inexpensive stainless steel or nickel steel alloy chosen for comparatively low thermal conductivity and tribological compatibility with the nickel steel valve seats. Airflow restriction in the ducts must minimize port pumping friction, and cam geometry must minimize valve throttling friction. The cylinder bore and piston rings will be cast of conventional materials to assure good lubricity and long life at minimal cost. Because the cylinder will be made of conventional materials which are thermally conductive, the combustion chamber will only be fully insulating when the piston is within 30 degrees of TDC. With the brief combustion reaction near TDC, combustion chamber temperatures will have dropped considerably by the time the cast iron cylinder bore is significantly exposed to combustion chamber gasses, minimizing heat energy loss.
While a steel containing 40% nickel is the preferred insulating material in the IPC engine prototype, it may not be the best in future applications. Discrete ceramic components may provide improved insulating performance, but are brittle and will require a significant development effort to reliably incorporate into this engine.
One popular discrete ceramic used in the adiabatic engine experiments of the early 1980s is called partially stabilized zirconia (PSZ), refer to SAE Technical Papers 820429 (1982) and 830318 (1983), whose abstracts are currently viewable at www.sae.org/technical/papers, and where the papers may be downloaded. These papers discuss internal combustion engine uses for PSZ. The thermal conductivity of two preferred PSZ ceramic compositions is approximately 2.0 W/m K.
PSZ ceramic was not sufficiently durable in the adiabatic engine experiments to become commercially applicable, though it performed remarkably well considering the severity of testing. It is expected PSZ will perform quite reliably at the lower average combustion chamber temperatures and milder thermal gradients within the IPC engine, but it must be incorporated in a manner which applies minimal tensile loading, preferring compressive loading where loads must exist.
Discrete ceramic insulators will not likely improve peak thermal efficiency of the IPC engine by more than a few percent over nickel steel, but the lower thermal conductivity of discrete ceramic insulators may improve overall thermal efficiency by permitting efficient operation of the engine at lower RPMs, such that the increased dwell of each combustion event at lower RPMs does not result in significantly increased heat energy absorption into the combustion chamber material. Discrete ceramic insulators may also improve CO exhaust emissions if it turns out the thermal conductivity of nickel steel remains sufficiently high to chill adjacent fuel.
Commercially available thin ceramic film coatings atop the nickel steel insulators may widen the engine’s RPM range and improve CO exhaust emissions in a manner much like discrete ceramic components. Since exhaust gas temperature in the IPC engine is low, many conventional emissions control devices will not function. Ceramic films strategically coating the nickel steel insulators may play a critical role in preventing the formation of CO exhaust emissions, without incurring the development requirements of discrete ceramic inserts. One potential complication with localized coating of the nickel steel insulators is there may be few commercially developed ceramic film coatings compatible with the low thermal expansion coefficient of the selected nickel steel alloy.
Powdered ceramic-metal composites may provide improved insulation performance over nickel steel and may also provide greater durability over discrete ceramics, but again, incorporation will require a great deal of development. Nickel steel, without ceramic film coatings, is currently preferred due to its low cost, ease of incorporation, and high durability. It is expected that crevice chamber pumping turbulence will adiabatically preheat critical nickel steel combustion surfaces sufficiently to prevent CO exhaust emissions.
As mentioned above, the combustion chamber predominantly insulates when the piston is within 30 crankshaft degrees of TDC. The combustion chamber partially insulates when the piston is farther than 30 degrees from TDC. As the piston travels from TDC toward BDC, and while the piston remains closer than 30 degrees to TDC, the heat generated by the combustion reaction is almost entirely dedicated to applying force to the crankshaft, finding minimal opportunity to route heat energy to the cooling system.
The combustion chamber switches from predominantly insulating to partially insulating when the piston drops below 030ATC, as a segment of thermally conductive cast iron cylinder bore starts to occupy a small portion of the combustion chamber's surface area. Combustion chamber gasses have adiabatically dropped in temperature by the time the thermally conductive cylinder bore surface becomes a significant percentage of the combustion chamber surface area, greatly reducing heat energy absorption into the cast iron cylinder.
Heat energy absorbed into the combustion chamber materials is not able to perform work on the crankshaft. It does not matter whether this absorbed heat is conducted into a cooling system or whether the absorbed heat returns back into the combustion chamber. Minimizing heat energy absorption into the combustion chamber material at the lowest manufacturing cost and lowest operating cost is a primary goal of a thermally insulating combustion chamber.
The thermally insulating segments of the combustion chamber exist to reduce heat energy absorption, thereby preserving heat energy for mechanical work. The thermally conductive segments of the combustion chamber exist in order to circumvent the significant tribological development requirements associated with using thermally insulating materials as wear surfaces.
Fig18 – Oil pump drive gear shown behind vibration damper. Oil pump not installed.




The thermally conductive cast iron cylinder bore cyclically forms a portion of the combustion chamber and will absorb a small portion of the heat of combustion. The average cyclic temperature of the cast iron cylinder bore remains below that which requires active cooling. On the other hand, the thermally insulating nickel steel portion of the combustion chamber will slowly absorb some of the heat of combustion and may need to transfer this heat away. The cooling method will be managed by ordinary oil circulation within the engine. The oil circulation system will assure all parts of the engine are lubricated as required, and all are kept at functional temperatures. Should the oil temperature climb to a designated upper limit, an external oil cooling circuit will activate. This external cooling circuit will possess a small radiator and blower fan, but the fan should not blow on the engine directly or the engine may overcool. When wind and cold weather will be present, the IPC engine may be best suited to operate in an enclosure without ambient venting.
Since the exhaust gas in the IPC engine is comparatively cool and dense, it may be more efficient to install an oil cooling coil within the exhaust plenum instead of an external oil cooling circuit with blower fan. A cooling coil inside the exhaust plenum may warm the oil more quickly on start-up and would keep the oil at a defined operating temperature. The operating temperature range of exhaust gas in the IPC engine is not yet known, and may be higher than permissible for lubricating oil, but this option will be retained for future consideration.
Fig19- Dry sump oil pan. Left cover houses 2 flywheels, 2 external dampers on right.




Combustion in the IPC engine is engineered for an unconventionally high heat release rate to assure complete combustion near TDC. The ordinary methods selected to achieve a high heat release rate are:
  1. Fuel-air charge fully mixed prior to ignition
  2. High compression ratio
  3. Significant chamber turbulence during ignition and combustion
  4. Fuel-lean stoichiometry selected to optimize rapid, clean combustion
In order to be able to select a preferred fuel-lean stoichiometry, the IPC engine tightly stratifies the combustion reaction within an optimally shaped volume at the centermost portion of the combustion chamber, assuring the speed, magnitude, and emissions level of the combustion reaction is optimal and consistent.
With a high heat release rate, the combustion reaction is initiated and consumed near TDC, achieving the combustion chamber’s temperature peak and pressure peak near TDC, thus allowing the gasses to begin adiabatically cooling and dropping in pressure right from the start of the expansion stroke. To prevent excessive temperature and pressure peaks in the combustion chamber at TDC, the maximum quantity of fuel in each combustion reaction must be comparatively small, limited in the IPC engine at full throttle to roughly one eighth the volume of fuel used in Otto and Diesel engines at full throttle. This limitation drives the low volumetric efficiency parameter of the IPC engine in order to prevent destructive detonation pressures sometimes associated with Otto cycle engines. Because temperature and pressure drops off early in the expansion stroke, the average temperature of combustion chamber gasses through an entire engine cycle in the IPC engine is comparatively cool. The expansion stroke is designed to maximize use of this limited pressure energy by defining the end of the expansion stroke at the point where cylinder pressure drops to a constructive value nearing an arbitrarily selected 1.3 ATM in which expansion force gains still exceed friction force losses.
In conventional Otto and Diesel cycle engines, the combustion flame must propagate well into the expansion stroke. While this is optimal for generating lots of power each engine cycle, the longer burn also retains significant heat in the combustion chamber late into the expansion stroke. This late burning fuel, for instance, fuel burning at 045 degrees ATC, can only expand at roughly a 6:1 ratio before being expelled to the exhaust, resulting in greater heat energy loss to the exhaust than if it combusted near TDC. Unfortunately, the large volume of fuel in Otto and Diesel engines cannot all be combusted at TDC without exceeding the pressure limits of the combustion chamber, so the Otto and Diesel engines reduced burn rate is necessary to achieve high volumetric efficiency.
The IPC engine borrows characteristics from Diesel, Otto, HCCI, and Atkinson cycle engines. Additionally, the IPC engine adds the selectively insulated combustion chamber described above, and incorporates a transitional chamber, called a crevice chamber, to assist combustion by stratifying fuel away from chilly crevices within the combustion chamber, constraining it to the center of the combustion chamber while generating turbulence which both premixes fuel with air and also accelerates the combustion reaction.
Specifically, the IPC engine borrows the unthrottled air intake and the direct fuel injection of a Diesel engine, but precisely injects fuel only from 035BTC until 025BTC (Figs20, 21). At these piston positions, the direct injector nozzles are aimed to inject fuel mass only into the piston pocket at the center of the combustion chamber. The crevice chamber’s pumping action directs fresh air toward the center of the combustion chamber to stratify the fuel while generating toroidal turbulence which mixes the injected fuel with air in the stratified central chamber. To sufficiently premix the fuel and air, the IPC engine must delay ignition until a few degrees before TDC, therefore requiring the spark ignition of an Otto engine. Combustion chamber pressure in the IPC engine is actively maintained slightly below the autoignition level, though occasional autoignition is quite acceptable if it occurs within a few degrees of TDC.
Fig20 – Spark plug and direct fuel injector concept. Spark plug body is nickel steel.




Like an HCCI engine, the IPC engine's combustion reaction is initiated and rapidly consumed near TDC. Unlike the HCCI engine, which uses compression ignition to reliably combust a dozen or so degrees before TDC, the IPC engine must use spark ignition to rapidly combust just a few degrees before TDC in order to provide the injected fuel sufficient time to turbulently mix with air in the piston pocket (Fig25), thus preventing soot emissions and assuring rapid combustion. Also, ignition just a few degrees before TDC exposes combustion chamber surfaces to peak temperatures for a marginally shorter duration of time before the expansion stroke begins, keeping the thermal insulators cooler and reducing bearing load duration, thus retaining more energy for mechanical work output.
The IPC engine also borrows the extended expansion stroke from the Atkinson engine to reduce pressure and heat energy levels in the exhaust duct, further reducing combustion chamber surface temperatures, reducing need for a cooling system.
The IPC engine operates with sufficient induction air volume to adiabatically pre-warm the pure air charge during compression to just below the auto-ignition temperature, readily promoting ignition when a spark is generated near TDC. This puts the dynamic compression ratio (DCR) close to 15:1, which will need to be continuously regulated to assure the pressure remains just below the autoignition level as conditions change (Figs22, 23, 24).
The 15:1 DCR ratio described here is independent of the 30:1 dynamic expansion ratio (DER) which will be described below. Note that the "DCR" will be referred to as the "compression ratio", and the "DER" will be referred to as the "expansion ratio".
It is unfortunate that internal combustion engines are usually designed with compression and expansion ratios being equal, as this mechanical simplicity perpetuates an erroneous notion that the exhaust duct must expel significant combustion energy in the form of heat and pressure. When an extended expansion ratio is coordinated with the rapid combustion reaction of HCCI and IPC engines, combustion energy can be most efficiently harnessed by the crankshaft, resulting in a cool exhaust that needs minimal muffling. The slow combustion reaction of Otto and Diesel engines is not as efficient at extracting energy from an extended expansion ratio.
Fig21 – Spark plug/fuel injector assembly is keyed to fit one way. Assy threads into head.


Fig22- Servo actuator shown at left cyl, servo link & camshaft lobes shown at right cyl.




If the IPC engine has a 100mm piston stroke, a 15:1 compression ratio, and a 30:1 expansion ratio, then the intake stroke occupies the first 50mm of piston travel after TDC and the compression stroke occupies the final 50mm of piston travel before TDC. Combustion chamber pressure will drop as low as 0.50 ^ 1.4 = 0.38 ATM in the period between the end of the intake stroke and the start of the compression stroke. Minimal duct flow is an important consideration in a low volumetric efficiency engine. A free flowing intake duct and crisp valve actuation will be used to minimize friction energy loss related to pumping air through the intake duct and throttling air past an actuating valve.
Fig23 – Conceptual adjustable valve timing, shown without camshafts or servo actuators.




The IPC engine must actively regulate the compression ratio to maximize cylinder pressure while preventing autoignition. The IPC engine will adjust the compression ratio by monitoring ignition reactivity and will use this data to continuously servo intake valve closure (IVC) timing. To detect ignition reactivity, a multi-spark algorithm used in the IPC engine initially provides a low energy spark a few degrees before TDC, followed microseconds later by a series of increasing spark energies. The spark event which triggers combustion will be detected by the engine controller. The energy level of the triggering spark will determine whether ignition occurred too easily, with difficulty, or properly, and the compression ratio will be servo-adjusted as indicated. Stratification of the fuel-air charge within the turbulent central combustion chamber of the IPC engine permits optimizing fuel-air “equivalence” ratios to assure precise control of ignition and combustion.
A fuel-air equivalence ratio other than 1.00 represents a fuel and oxidizer ratio deviation from stoichiometric. Stoichiometric fuel-air, as often found in Otto and Diesel engines at full throttle, has a 1.00 equivalence ratio. The HCCI engine at full throttle must typically have a fuel-lean 0.25 equivalence ratio, since all combustion occurs near TDC and higher equivalence ratios may cause piston knocking conditions. The HCCI engine at full throttle consumes roughly one quarter the fuel during each combustion event as Otto and Diesel engines at full throttle.
For a given cylinder displacement, the IPC engine at full throttle consumes roughly half the fuel each combustion event of similarly displaced HCCI engines, and therefore consumes an eighth of the fuel of similarly displaced Otto and Diesel engines at full throttle. This is because the IPC engine only uses half the piston stroke for compression, and therefore the combustion chamber volume of the IPC engine at TDC is roughly half the size of HCCI, Otto, and Diesel engines of similar displacement at TDC. While the IPC engine uses only half the piston stroke for compression, it uses the full piston stroke for expansion.
The HCCI engine at part throttle has an equivalence ratio that drops below 0.20, but this is sufficiently lean to slow the combustion reaction rate, which also tends to generate unsatisfactory levels of exhaust pollution. This lean condition is a primary area of research with HCCI engine prototypes today. The IPC engine is constrained by the same cylinder pressure limitations of HCCI, Otto, and Diesel engines, however the IPC engine uniquely permits tailoring the volume of the stratified central combustion chamber to permit selecting fuel-air equivalence ratios which optimize both combustion speed and clean exhaust emissions.
The IPC engine’s stratified combustion chamber retains pure air at the periphery of the combustion chamber while tightly stratifying fuel to the turbulent central combustion chamber, permitting uniform equivalence ratios of 0.50 throughout the central chamber while retaining a pure air equivalence ratio of 0.00 in the periphery. Equivalence ratios in the 0.50 range are sufficiently oxygen rich to assure all fuel is oxidized, and the comparatively high percentage of fuel causes the reaction to hotly progress at the fastest rate, the combustion “rate” defining the highest permissible RPM which supports complete combustion at low emissions, and therefore at the highest permissible horsepower and thermal efficiency. It should be noted that the IPC engine operates most efficiently in a narrow RPM band near the highest permissible RPM, since operating significantly below the highest permissible RPMs increases dwell time for heat soak of the combustion energy into the combustion chamber material, since the nickel steel thermal insulation material selected for the IPC engine is comparatively efficient, yet it does absorb some heat energy. The briefer heat soak allows more energy from each combustion event to reach the crankshaft.
The fuel air equivalence ratio of the central combustion chamber of the IPC engine is 0.50 at full throttle. The IPC engine can reduce the equivalence ratio to 0.30 in the central combustion chamber at low throttle and still retain a sufficiently rapid combustion reaction to achieve low exhaust emissions, eliminating the need for emissions control components. The IPC engine can actually handle equivalence ratios slightly below 0.20, due primarily to the significant turbulent kinetic energy within the central combustion chamber at time of ignition, but dropping the equivalence ratio significantly below 0.20 in the IPC engine will generate an unsatisfactorily slow combustion rate which results in an increase in exhaust emissions. The IPC engine is presently modeled to remain well above this lower boundary as an emissions safety factor. At present, I lack data which will define an upper boundary which generates rapid, clean combustion, so I will arbitrarily use the 0.50 equivalence ratio as an optimal upper bound “placeholder” until applicable data emerges. The IPC engine will be tasked to operate only at equivalence ratios which are high in combustion speed as well as low in pollution emissions. In summary: It is expected the IPC engine will operate effectively at an equivalence ratio range from 0.30 to 0.50.
Intake valve closure (IVC) may be driven by a servo-controlled camshaft roller follower to assure proper air volume is induced to achieve sufficient compression pressure under all engine conditions. Intake valve opening (IVO) timing can be driven by the same camshaft, but will use a pair of roller followers without variable timing. Two roller followers will be used for IVO due to the greater loads associated with opening multiple valves compared with closing them. Since intake valve actuation will occur during a brief 90 crankshaft degrees, it may be necessary to employ a unique roller camshaft with comparatively large base circle to assure responsive valve actuation which minimizes valve throttling friction losses. This “larger than ordinary” intake camshaft base circle suggests a need for an intake camshaft completely separate from the exhaust.
It should be noted that the IPC engine concept presented here shows a most primitive valve actuation mechanism intended only to provide functional definition of the engine concept. The valve actuation mechanism is impractical in this form, and exists as a “placeholder” design, intended to be replaced by practical valve actuation schemes as time permits.
There may be cost benefit to individually servoing IVC timing at each cylinder to reduce the need to blueprint all cylinders to exacting tolerances, and to compensate for variations caused by normal wear in the engine, however it also appears individual cylinder adjustments may not be necessary with the highly reactive ignition and combustion processes allowed by the IPC engine's intrinsic stratification efficiencies. Still, the present IPC concept servo-adjusts the IVC timing of each cylinder individually as a safety factor, using linear actuators with a 20mm stroke, but bulk adjustment of the intake valves at all four cylinders using a single actuator may remain a more cost effective direction as information becomes available.
If the engine has a 100mm piston stroke, the expansion stroke occupies 100mm of piston travel after TDC, and the exhaust stroke occupies 100mm of piston travel prior to TDC.
As mentioned above, the compression ratio of the IPC engine is independent of the expansion ratio, and the expansion ratio will be about 30:1 to minimize heat energy loss to exhaust duct, much the way an Atkinson engine minimizes exhaust energy loss. The 30:1 expansion ratio is based on a peak pressure capacity of 150 ATM at TDC arbitrarily assumed to be limits which prevent the formation of oxides of nitrogen pollutants, and on the prevalence of predominantly diatomic gasses of the fuel-lean combusted charge obeying, to a first order approximation, the 150 ATM / (30 ^ 1.4) = 1.3 ATM equation. Note that prevention of oxide of nitrogen exhaust emissions drives the 150 ATM specification at TDC, and mechanical friction amplitude drives any deviation from the 1.3 ATM specification at BDC, and they will be adjusted as required.
The unconventionally large expansion ratio is chosen to extract virtually all useable heat and pressure from the combustion chamber before the exhaust valve opens, resulting in a comparatively cool and quiet exhaust stroke with minimal exhaust duct flow velocity. Additional thermal efficiency may be gained at low and moderate throttle positions by servoing the exhaust valve open (EVO) timing event until well past BDC to effectively reduce the 30:1 expansion ratio as low as 20:1 to prevent port pumping losses related to the smaller combustion reaction. The present IPC engine shows the conceptual mechanism which servo-adjusts EVO timing.
Fig24 – Conceptual valve timing adjusters shown, linear servo actuators not shown.




Incorporation of a transitional chamber, called a crevice chamber, at the outer perimeter of the combustion chamber of the IPC engine assists complete combustion (Fig25). The crevice chamber forms as the piston travels within 40 degrees of TDC, effectively creating a central combustion chamber that accepts injected fuel, a peripheral annular crevice chamber that rejects injected fuel, and an intermediate annular passage that communicates between the two chambers. The crevice chamber contains the combustion chamber’s crevice-containing features, such as piston rings, head gasket, and valve seats. Note that the volume of the crevice chamber approaches zero at TDC, whereas the volume of the central combustion chamber approaches a finite value at TDC, creating an effective pure air pump directed from the crevice chamber toward the central chamber in the last 40 degrees before TDC.
The annular passage, or “backfill passage”, between the two chambers is sized at TDC with roughly three times the air volume of the crevice chamber at TDC, and is shaped with reduced surface area and without crevices in order to support combustion when the combustion reaction ignites and then backflows out of the central combustion chamber. This backfill passage (Fig26) is designed to induce a laminar flow between the central combustion chamber and the crevice chamber at time of combustion, and the backfill passage volume assures only pure air backfills into the chilled creviced areas of the combustion chamber, retaining all fuel in the central combustion chamber and backfill passage. When combusting gasses begin to expand into the backfill passage, it does not significantly mix with the pure air in the backfill area, but rather it pushes much of the backfill pure air into the crevice chamber, therefore the combusting gasses retain a consistently fast reacting fuel-air equivalence ratio throughout the entire stratified combustion reaction. Note that, since combustion begins and ends at TDC, there is not a second opportunity to combust residual fuel that does not ignite at TDC. All fuel must combust at TDC, and all areas of the central combustion chamber and backfill passage must support combustion. Creviced features, such as valve seats and spark plug insulation recesses, are not permissible in the combustion area if exhaust emissions are to be low.
At present, the central combustion chamber at TDC is sized to be the same volume as the crevice chamber plus backfill passage at TDC. This is predicated on data which indicates 0.25 is the maximum equivalence ratio of an HCCI engine, and on the arbitrary assumption that an equivalence ratio of 0.50 is the highest that will combust rapidly and with low emissions. An investigation continues into equivalence ratio boundaries. If it turns out that equivalence ratios such as 0.65 combust appropriately in the central combustion chamber, then it can be expected these higher equivalence ratios will be targeted and the central combustion chamber may decrease in volume with respect to the combined crevice chamber and backfill passage volumes.
Fig25 – Transitional crevice chamber & annular passage are formed at 40 degrees BTC.




The IPC engine is a reciprocating piston, spark ignition, 4-stroke engine, but it is sufficiently unconventional that stepping through a full engine cycle may help with understanding. The IPC engine possesses twelve stages of operation named, 1) Intake, 2) Vacuum, 3) Rebound, 4) Compression, 5) Injection, 6) Turbulence, 7) Ignition, 8) Combustion, 9) Expansion, 10) Vacuum, 11) Rebound, and 12) Exhaust. We will step through all of these stages to get a sequential picture of how this engine operates:
359BTC: Intake valve opens, drawing in fresh unthrottled air, same as a Diesel. 270BTC: Intake valve closes as piston falls 50% of stroke distance from TDC. 269BTC: Cylinder begins pulling a vacuum as piston continues downward. 180BTC: Combustion chamber vacuum peaks at 0.50 ^ 1.4 = 0.38 ATM air pressure. 179BTC: Piston elastically rebounds off vacuum, being efficiently pulled upward. 090BTC: Vacuum rebound ends, compression of fresh air begins. 089BTC: Fresh air begins adiabatically heating. 040BTC: Combustion chamber divides into central chamber and crevice chamber. 035BTC: Fuel is direct injected toward pocket at center of piston. 034BTC: Crevice chamber pumps fresh air toward piston pocket, constraining fuel. 030BTC: Combustion chamber becomes predominantly thermally insulating. 025BTC: Crevice chamber generates toroidal vortex in piston pocket, mixing fuel and air. 005BTC: Fuel and air is constrained to, and homogenously mixed in, central chamber. 004BTC: Spark ignites fuel and air mixture, fuel-lean combustion is violent. 000TDC: Cylinder pressure increases, combusting gas expands into backfill passage. 001ATC: Laminar flow within backfill passage returns pure air to crevice chamber. 004ATC: Fuel-lean combustion is violent and complete. Reaction extinguishes. 005ATC: Peak cylinder temperature and pressure begins dropping adiabatically. 029ATC: Combusted gasses are adiabatically cooling in combustion chamber. 030ATC: Combustion chamber first exposes thermally conductive cylinder bore. 040ATC: Central and crevice chambers transform back into single combustion chamber. 135ATC: Combustion chamber starts pulling a vacuum (low throttle only). 165ATC: Combustion chamber starts pulling a vacuum (mid throttle only). 179ATC: Combustion chamber pressure drops to 1.3ATM (full throttle only). 180ATC: Combustion chamber pressure or vacuum depends on throttle position. 181ATC: Expansion stroke ends, exhaust valve opens (full throttle only). 195ATC: Combustion chamber vacuum ends, exhaust valve opens (mid throttle only). 225ATC: Combustion chamber vacuum ends, exhaust valve opens (low throttle only). 359ATC: Exhaust valve closes. Fig26 – Backfill passage has three times the volume of the crevice chamber at TDC.




This first iteration of a complete IPC combustion chamber shows eight tiny valves per cylinder surrounding a circular central chamber (Fig27). This is the theoretically simplest construction that meets the needs of the IPC engine, though there may be physically simpler constructions found during development which will reduce cost. For now, the four tiny intake valves and four tiny exhaust valves exist at the perimeter, and the central chamber will remain circular and symmetrically centered. Individual valves can be deleted as testing determines they are not needed.
Fig27 – 8 valves per cylinder is the simplest theoretical arrangement for the IPC engine.




Since the combustion reaction rate slows as the fuel-air equivalence ratio drops below 0.20, and since exhaust pollutants become a significant issue when the reaction rate slows, the IPC cannot throttle fuel outside the 0.30 to 0.50 equivalence ratio boundaries without generating significant levels of exhaust pollutants. It is also not thermally efficient to run the IPC engine at low RPMs, as lower speeds increase heat soak dwell time of the combustion chamber materials during each combustion event. These two conditions restrict the horsepower band of the IPC engine to a fairly narrow range. It is necessary to selectively shut down individual cylinders in the IPC engine when a wider horsepower range is required. To permit efficient shut-down of selected cylinders, greater EVO valve timing adjustment range becomes necessary on the deactivated cylinders to reduce parasitic energy losses, mandating that the variable EVO valve timing be individually adjustable at each cylinder (as shown), and requiring the exhaust camshaft base circle diameter be increased to match that of the intake camshaft.
As noted above, significant issue in the IPC engine is heat soak of the combustion chamber, since nickel steel is not a perfect thermal insulator. It is a goal that combustion heat energy of each reaction be present in the combustion chamber for the shortest possible duration of time each engine cycle. For this reason, the IPC engine will likely operate over a rather narrow RPM range, defined as an RPM compromise which provides sufficient combustion reaction time near TDC to fully combust the fuel-air mixture to prevent exhaust emissions, while rapidly expanding the combusted gas to minimally expose combustion chamber materials to the heat soak dwell of each combustion event. It is expected the IPC engine may best operate within the arbitrarily selected range of 3200 to 4000 RPM. Operating above 4000 RPM will generate polluting emissions, operating below 3200 RPM will measurably decrease thermal efficiency.
A higher and wider RPM range may be realized when data becomes available to analyze combustion rates as they relate to exhaust emissions at higher RPMs. At present, my only available data is for lower RPMs, as found in HCCI technical papers, and I’ve done some extrapolation to come up with the arbitrary 4000 RPM limit for the IPC engine. Note that most emissions control components will not operate efficiently with the IPC engine due to the cool exhaust, so the IPC engine concept must be designed to combust cleanly without emissions controls to maximize economic merit.
Since normally aspirated 3.2 liter Otto and Diesel engines will typically generate about 200 horsepower at 4000 RPM, it can be assumed the IPC would generate 0.125 * 200=25 HP, if the energy efficiency was the same between engines. Since it is arbitrarily expected the IPC engine will be twice as energy efficient as the peak efficiencies of similarly displaced Otto and Diesel engines, it is expected the 3.2 liter IPC engine will produce 50 HP at 4000 RPM.
The narrow operating range of the IPC engine, regarding RPM and equivalence ratio, suggests the optimal horsepower range of an IPC engine is quite a narrow band. If the IPC engine puts out a maximum sustainable 50 HP, the minimum horsepower will be (.30/.50) * (3200/4000) * 50 = 24 HP, with a dynamic operating range for the IPC engine from 24 HP to 50 HP. As noted above, greater dynamic range requires selectively deactivating individual cylinders, such that 4 cylinders, 3 cylinders, 2 cylinders, or 1 cylinder of the engine may be selectively active at a given time. Being able to selectively shut down cylinders generates the four unique dynamic horsepower ranges of 6 HP to 12 HP, 12 HP to 25 HP, 18 HP to 37 HP, and 24 HP to 50 HP. This means the IPC engine can operate efficiently from 6 HP to 50 HP. This is a highly functional operating range, assuming the deactivated cylinders are constructed such that they do not consume significant parasitic energy.
It may be desirable to operate the IPC engine below the 6 HP level, but slowing the RPM below 3200 RPM will increase dwell time for heat soak of each combustion event. It can be expected that RPMs below 3200 will cause the thermal efficiency to decrease slightly. Fortunately, this will reduce thermal efficiency at a time when only one cylinder is consuming fuel, and it is being consumed at the lowest equivalence ratio, so minimal fuel is being consumed, allowing the engine to retain reasonable thermal efficiency. It can be expected that the IPC engine will run effectively as low as 1000 RPM, and this slower RPM may also permit equivalence ratios below 0.30 sufficient burn time to combust cleanly, allowing the engine a bonus thermal efficiency at low RPMs, thus defining an effective dynamic operating range for the IPC engine from 2 HP to 50 HP.
One proposal describing the full engine cycle of a deactivated cylinder in the 4-stroke IPC engine:
359BTC: Intake valve opens, drawing in fresh unthrottled air, same as a Diesel. 270BTC: Intake valve closes as piston falls 50% of stroke distance from TDC. 269BTC: Cylinder begins pulling a vacuum as piston continues downward. 180BTC: Combustion chamber vacuum peaks at 0.50 ^ 1.4 = 0.38 ATM air pressure. 179BTC: Piston elastically rebounds off vacuum, being efficiently pulled upward. 090BTC: Vacuum rebound ends, compression of fresh air begins. 040BTC: Combustion chamber divides into central chamber and crevice chamber. 034BTC: Crevice chamber pumps fresh air toward piston pocket. 025BTC: Crevice chamber generates toroidal vortex in piston pocket, generating friction. 000TDC: Cylinder pressure peaks. 040ATC: Central and crevice chambers transform back into single combustion chamber. 090ATC: Combustion chamber starts pulling a vacuum 180ATC: Combustion chamber vacuum peaks again at 0.38 ATM. 181ATC: Piston elastically rebounds off vacuum, being efficiently pulled upward. 270ATC: Vacuum rebound ends, exhaust valve opens. 271ATC: Expulsion of pure air into exhaust port begins. 359ATC: Exhaust valve closes.
Note that the deactivated cylinder algorithm above contains slight inefficiencies which generate port flow friction losses without the benefit of work output. More study may reveal construction alternatives which will reduce this parasitic loss. An example would be if crankcase vacuum is present to reduce crankcase turbulence, this vacuum source may be applied to the intake and exhaust ducts of the deactivated cylinders. Applying the vacuum to the ducts may also permit simplification of the shown EVC timing mechanisms.
An overpressure bypass valve is incorporated into this early concept engine (Figs 24, 26, 28), but will likely be deleted when it is calculated there will be no progressive physical damage done to this nickel/steel variation of the IPC when stoichiometric combustion reactions accidentally occur. Future ceramic variations may need to reemploy this protection feature.
Fig28 – Port for overpressure bypass valve is shown, likely not needed in IPC engine.




Friction reduction in the IPC Engine

Mechanical friction and windage friction are significant issues in a low volumetric efficiency engine like the IPC. The following are some of the friction reducing methods being considered for this engine.
Twin gear-synchronized counter-rotating crankshafts eliminate piston side thrust forces, eliminating skirt thrust friction and piston slap. The piston rings are more stable in their grooves when side thrust is eliminated. Twin crankshafts also eliminate the need for crankshaft counterweights (Fig30), allowing two crankshafts to weigh about the same as a single counterweighted crankshaft. Note that twin crankshafts in the IPC engine each have 90mm strokes to achieve the 100mm piston stroke, reducing crankshaft mass and increasing stiffness further. This construction leads to a BDC piston position only 157 degrees after the TDC piston position, a characteristic which can be used to further reduce hot time on combustion chamber surfaces if the crankshaft rotation direction is selected for this benefit. The quicker expansion cycle described also speeds the intake cycle valve actuation, such that the IPC engine’s 90 degree intake cycle when using a single crankshaft engine consumes only about 80 crankshaft degrees when using dual crankshafts, forcing the camshaft lobes to actuate the valve very quickly, and likely requiring a larger camshaft base circle. Similarly, the slower compression stroke described reduces turbulence velocity just prior to combustion in the combustion chamber. These issues must all be considered.
Presently, it is only necessary to recognize that the expansion stroke and the compression stroke are uniquely asymmetrical with this twin crankshaft geometry, as are the intake and exhaust strokes, and the crankshaft rotation direction can be selected later (Fig29), based on priorities. Note that the asymmetry also means balance is not intrinsically perfect (all four pistons do travel the same direction at the same time twice each crank revolution), but it presently looks to be an insignificant issue resulting in a smooth running engine, with counterbalance mechanisms optionally available to further improve smoothness. The dual-crankshaft's synchronizing gears are helical so they operate smoothly and silently, and the pitch diameter is large enough to minimize tooth sliding friction under load. The gear tooth helix angle is reduced by selecting a wide 25mm tooth, keeping loads on the crankshaft thrust bearings low. There may be insufficient benefit to using twin crankshafts over a single crankshaft, but twin crankshafts are presently implemented in the IPC engine as a research tool.
Fig29 – Dry sump oil pan design is determined by crankshaft rotation direction.




With the elimination of piston side thrust forces, the piston skirt can be sculpted to reduce contact surface area with the cylinder bore without concern for skirt wear, resulting in reduced piston skirt viscous oil friction losses. Two circumferential contact bands and ten vertical contact stripes are shown, but a subset of this configuration may be selected.
Since the insulated pulse engine has high cylinder pressures for a very short percentage of crankshaft rotation, gas ported low-tension piston rings reduce piston sliding friction through the bulk of the piston travel while sealing tightly when pressure is high. The IPC engine’s combustion chamber is transitionally stratified and permits only pure filtered air to enter the piston’s gas ports, preventing localized HC exhaust emissions and also preventing gas port clogging. Gas ported piston rings may be calibrated to apply too much pressure to the cylinder walls when pressure is high, possibly causing excessive piston sliding friction and excessive cylinder wear at the very top of the bore, so the details will need to be researched. Presently, there are no production engines which use gas ported pistons, only racing engines, but production engines differ from the IPC engine in that they all permit fuel to enter the gas ports, and this results in HC pollution issues.
Fig30 – Crankshaft of inline 4-cylinder twin crankshaft engine needs no counterweights.




With the short duration of high pressures, and lack of piston side thrust and side slap, it may be possible to reduce the number of compression rings from two to one, reducing piston sliding friction further. This may require incorporating a commercial gapless ring technology, if gapless rings support gas porting. A lightweight ductile iron piston may be required to remove the second compression ring, as a ductile piston can more snugly fit the iron cylinder bore and allow tighter ring groove clearances than aluminum pistons. Aluminum pistons, however, run cool in IPC engine and may be fitted quite snugly without issue, and may be easier to manufacture with the nickel steel insert. The nickel steel insert may extend downward to comprise one or more piston ring grooves, but it presently seems there are no tribological, thermal conductivity, or thermal expansion bonuses in doing so, so it will not be investigated at this time. This is an item to research, and will be discussed with ring manufacturers as time allows.
The shown intake IVC timing mechanism is a first generation concept, the shown exhaust EVO mechanism is a second generation design. Both theoretically function but neither are suitable for fabrication at the present revisions. As shown, both the intake and the exhaust camshaft follower mechanisms require 20mm linear servo actuators to function. In a few more generations, the valve timing adjust mechanisms are intended to be weight reduced, the valve tip followers will be hydraulically self-adjusting, and the four IVC actuators may be replaced by a common IVC actuator and linkage to gang-actuate the intake valves of all cylinders. More likely, these mechanisms will soon be replaced by practical designs which currently exist in production engines.
Pushrod-actuated valve train mechanisms often have components which simultaneously reciprocate in opposing directions, aiding the reduction of engine vibration. The IPC engine shown does not presently have this opposing action, and a supplemental mechanism may be employed which actuates by auxiliary lobes on the camshaft ends in order to artificially generate the optimal counterbalance action of both the valve train and pistons to cancel all engine vibration for a slight net gain in flywheel horsepower.
Fig31 – Crankcase has extra space to minimize air entrapment energy losses.




Intake port friction reduction: Intake air in the IPC engine is unthrottled, as it is in Diesel engines. Throttled intake air, as found in Otto engines, consumes more pumping energy than unthrottled intake air. Similarly, for a predefined volume of air, rapid crisp intake valve actuation “throttles” less than a more gradual intake valve actuation, therefore crisp actuation requires less pumping energy for the same volume of induced air. Crisp valve actuation, if sufficiently durable, is sought for the IPC engine’s intake port. It is expected the four tiny intake valves will flow with low resistance.
Exhaust port friction reduction: There is minimal opportunity for “tuning” the exhaust with headers, since there is so little energy in the expelled exhaust. The exhaust ports are simply made short and free-flowing, and dump into an aluminum exhaust plenum that directs to a low restriction outlet pipe. Since the exhaust is comparatively cool, and since the exhaust flow velocity is low and consumes little energy, it is expected the IPC engine will need little or no muffling. It is expected the four tiny exhaust valves will flow with low resistance.
Fig32 – Bulkhead between cyls 1&2 and 3&4 are vented to reduce crankcase turbulence.




Combustion chamber turbulence friction reduction: Turbulence is preferably generated only from time of fuel injection until the instant combustion extinguishes, to optimally pre-mix fuel and air and promote rapid combustion. Turbulence generated significantly prior to fuel injection, such as the conventional tumble or swirl options, is wasted vortex energy. Turbulence generated after the combustion process extinguishes is also wasted vortex energy. The IPC engine’s construction targets this goal as close as possible. If insufficient turbulence is generated to fully mix the fuel and air, combustion friendly protrusions or fins may be added to the piston pocket or to the spark plug insert to increase turbulent kinetic energy (TKE).
Crankcase turbulence friction reduction: Block bulkhead ports between pistons 1&2 and pistons 3&4 relieve an air pressure transient in crankcase caused by opposite piston motion. The present crankcase volume is oversized to minimize air entrapment issues which generate windage losses, but this oversize can be undesirable. A crankcase vacuum pump may alternately be employed, particularly at higher RPMs, to reduce crankcase windage friction, permitting reduced crankcase volume, though oiling circuitry will need to be adapted to work within the vacuum gradient that is generated within the engine. For instance, it may be necessary to locate the pump in the base of the oil pan to gravity-feed oil into the pump. Crankcase pressure when the vacuum is applied is presently targeted at 0.25 ATM, but further research may find a more optimal value.
Fig33 – IPC engine block is cast iron to match temp-coeff of crank and connecting rods.




Unfinished topics to be added to this draft

The engine block is cast of iron to best match the thermal expansion coefficient of the nodular iron crankshaft and steel connecting rod assemblies (Fig33). This thermal stability allows the crevice chamber to be very small at TDC without concern for the pistons contacting the head or valves at temperature extremes or during temperature transients. Aluminum may replace the cast iron block when operating temperatures, transients, and chamber volume tolerances are characterized.
The items likely attached to the accessory pulley of the IPC engine may be a crankcase vacuum pump, air conditioner pump, and power steering pump. The accessory attachment bosses will likely be placed onto the cylinder block, possibly allowing the end covers and oil pan to be molded of plastic rather than cast of aluminum. Timing belt will be driven by primary crankshaft, accessories by secondary crankshaft to reduce energy transfer through the crankshaft gear teeth. Primary crankshaft is the one selected to drive the PTO flange, and is usually the crankshaft that will rotate the flange in the conventional engine direction.
Fig34 – Connecting rod assembly consists of two steel rods and a steel center link.




Four-bolt powdered metal connecting rods are shown (Fig34), but are only a design exercise, not intended to suggest a best method. The engine block main caps (Fig35) have a uniquely flexible bolt tab arrangement that is only a concept, not intended to suggest a best method. The adjustable valve actuators shown are presently first generation concepts (intake side) and second generation concepts (exhaust side). Further research is assumed necessary for many of the unusual engine features, and some are only intended as placeholders, to be replaced by more applicable mechanisms as the design matures beyond this first generation complete engine assembly.
It is a future goal to upgrade the shown side-by-side spark plug and fuel injector assembly. The goal is a concentric arrangement, with the injector’s tip perfectly centered in the combustion chamber and also acting as the spark anode, with fuel nozzles aimed to miss the electrode. A ceramic insulator supports the injector concentrically. Ceramic seals, combustion pressures, and high voltage issues are sufficiently complex that this is a future contemplation. For now, the goal is to integrate a mildly modified commercial injector with a custom spark plug module made from a disassembled spark plug. It is expected the ignition coils in the IPC engine will mount atop the valve cover at each cylinder. The valve covers have not yet been designed.
Spark plugs may get streamlined anodes which taper flush with a similarly tapered ceramic insulator in order to reduce crevice volume. Similarly, the spark plug electrode may become an airfoil to eliminate crevice volume. The ceramic may taper to match the spark plug body contour, eliminating another crevice. The direction and flow of actual turbulence in the combustion chamber will logically influence these shapes. Fuel injector nozzles may similarly take on this creviceless aero shape, particularly if it becomes concentrically integrated with the spark function.
The shown connecting rods and center link are heavy when compared to a single crankshaft engine. This seems a fair trade, as the greater mass will generate a slight extra bit of bearing and turbulence friction, but this is more than offset by the elimination of piston skirt side thrust forces which eliminates skirt side thrust friction and allows reduced piston skirt contact area to the cylinder bore, reducing oil viscosity friction at the skirt as well.
Unfortunately, the IPC engine is an “interference engine” by design. The shown timing belt will eventually need to be replaced by a timing chain, since timing chains typically need no replacement during the service life of an engine. This will need to be addressed.
Fig35 – Main caps are a conceptual 4-bolt design which reduce bulkhead flex under load.




A 2-stroke IPC engine construction.

A conceptual 3.2 liter in-line 4-cylinder 2-stroke IPC engine will now be considered. The 2-stroke IPC engine's combustion process appears efficient and clean. The 2-stroke and 4-stroke variations of the IPC engine are sufficiently distinct that both engine constructions are worth investigating. The 2-stroke IPC engine is now being modeled in CAD, and images will be presented here as they become available.
The intake ducts of the 2-stroke IPC engine are located at the base of the cylinder bore, as is commonly done in 2-stroke engines. Since the IPC engine is expected to apply a vacuum to the crankcase to reduce crankcase turbulence energy losses at higher RPMs, the IPC 2-stroke may not be able to use commonly employed methods to induct fresh air into the combustion chamber. There will likely be a need for the addition of an air pump on the 2-stroke IPC engine which provides a volume of fresh filtered air each combustion event. This discrete pump would be driven by the accessory belt of the engine, and would also assure the inducted air was free of oil, since crankcase sourced air can become saturated with crankcase oil mist. Piston skirt design would need to address intake port sealing when the piston is near TDC to assure fresh air remains free of crankcase oil, but the piston would otherwise be similar to the 4-stroke version. If there is a complication with the air pump, it might be related to overpumping energy losses when individual cylinders are deactivated at reduced power levels. Possible use of the crankcase as four discrete intake air pumps will be considered, and may present an interesting construction which also eliminates need for a discrete vacuum pump at the accessory belt.
The valves in the 2-stroke cylinder head are used only for exhaust. Despite elimination of the intake valves, the construction of the 2-stroke head remains quite similar to that of the 4-stroke head. For a given bore and stroke, which is 100mm bore X 100mm stroke for the IPC engine concepts being presented, the 2-stroke combustion chamber volume will have to be reduced slightly since the new intake port location consumes a portion of the piston stroke. The reduced combustion chamber volume permits larger valve head diameters than found on the 4-stroke IPC engine. At the present time, the 2-stroke concept will employ eight exhaust valves per cylinder.
Since this is a 2-stroke engine, the two exhaust camshafts will rotate at the same velocity as the crankshaft. This eliminates the large camshaft base needed in the 4-stroke IPC engine to achieve short duration valve actuations. Cam lobes can be quite conventional in the 2-stroke IPC engine. This will also free up territory for mounting the spark/injector assembly. The present 2-stroke concept employs a pair of exhaust camshafts in the same location as the 4-stroke IPC engine, but this may be reduced to a single camshaft as the mechanical CAD model develops.
The 2-stroke IPC engine retains the adjustable 15:1 compression ratio of the 4-stroke IPC engine, with the variable compression ratio determined by exhaust valve closure timing. Also like the 4-stroke IPC engine, the expansion stroke of the 2-stroke IPC engine remains a nominal 30:1 to assure high thermal efficiency. Since it appears deactivation of individual cylinders is the most efficient way to achieve reduced horsepower levels, variable timing for EVO will retain individual actuators at each cylinder, while variable timing for EVC may only require a single actuator which gang-activates all four cylinders at once.
In the present 2-stroke concept, nominal exhaust valve closure occurs when the piston rises to a position 38mm below TDC to achieve a 15:1 compression ratio with the slightly reduced 2-stroke combustion chamber volume. This puts the nominal exhaust valve opening when the piston falls to a position 75mm below TDC. This leaves the lowest 25mm of piston travel only for exhaust and intake functions, not for compression or expansion. The dwell time created by this 25mm lowers the average combustion chamber surface temperatures, significant since the 2-stroke has a combustion event every crankshaft revolution. The 25mm specification may be increased if average combustion chamber temperatures are too high, or the 25mm may be reduced if the combustion chamber materials can sustain higher average temperatures. To increase the 25mm specification in a 100mm stroke combustion chamber, the 75mm specification would need to be reduced, as would the 38mm specification.
Specifically, this 25mm defines the point at which the intake ports in the cylinder bore first become a segment of the combustion chamber, and the exhaust port is nominally timed to open just about the time this intake port first shows up. The intake port is slightly pressurized with reference to the gasses within the combustion chamber, so a productive upward laminar combustion chamber flow occurs with the intake and exhaust cycles beginning simultaneously. Fresh pressurized air enters the bottom of the combustion chamber while exhaust is pushed out the top of the combustion chamber.
As mentioned, the piston drops to reveal the intake port at roughly the same time the exhaust valve opens to begin the exhaust stroke, and the lower portion of the combustion chamber gradually fills with slightly pressurized fresh air to push the spent chamber gasses upward and out the opening exhaust valves at the top of the cylinder. As the piston reaches BDC to fully reveal the intake ports, and then moves upward, the intake port begins to close while the piston begins to push the combusted gasses at the top of the combustion chamber out the valves. When the intake ports becomes fully covered by the rising piston at 75mm before TDC, primarily fresh filtered air is found at the bottom and middle of the combustion chamber, while spent combustion chamber gasses stratified at the top of the chamber are being pushed out the exhaust valves by the rising piston. The piston continues pushing these upper chamber gasses out until the exhaust valves close and compression begins. The combusted gases of the IPE engine are inherently oxygen rich because of the fuel-lean burn, so they retain oxidizer value useable in future combustion cycles. Because of this, all combusted gasses in the IPC engine need not be purged each exhaust cycle, though nearly all are exhausted, making the 2-cycle IPC engine potentially quite efficient.
Exhaust emission characteristics of the 2-stroke IPC engine are likely as clean as the 4-stroke IPC engine, and the volumetric efficiency may be higher than the 4-stroke IPC engine, assuming temperatures and pressures which begin to form nitrogen emissions are the limiting parameter of the 4-stroke IPC engine. It makes sense that insulator temperatures stand a greater chance of being the limiting parameter of the 2-stroke IPC engine. There is not necessarily a thermal efficiency penalty in moving toward a cooling system of greater capacity in a 2-stroke if required to improve volumetric efficiency over a 4-stroke IPC engine, since the hotter "skin" temperature of the 2-stroke combustion chamber may absorb less heat energy from each combustion cycle. Since it looks as though both the 2-stroke and 4-stroke IPC engines may be limited to a 4000 RPM redline due to combustion reaction time requirements, the 2-stroke IPC engine may provide a full 50% volumetric efficiency improvement over a 4-stroke IPC engine. The result may be an IPC engine assembly which provides notably lower manufacturing cost and operating cost over similarly powered Otto and Diesel engine assemblies.
The 2-stroke IPC engine cycle sequence is as follows (degrees are approximate):
120BTC: Intake ports in cylinder bore become blocked by piston insulator. 119BTC: Lower portion of combustion chamber contains mostly fresh filtered air. 118BTC: Upper portion of combustion chamber contains mostly O2-rich exhaust. 117BTC: Piston pushes exhaust stratified in upper chamber out the exhaust valves. 070BTC: Exhaust valve closes, compression of fresh air and some exhaust begins. 069BTC: Fresh air begins adiabatically heating. 040BTC: Combustion chamber divides into central chamber and crevice chamber. 035BTC: Fuel is direct injected toward pocket at center of piston. 034BTC: Crevice chamber pumps fresh air toward piston pocket, constraining fuel. 030BTC: Combustion chamber becomes predominantly thermally insulating. 025BTC: Crevice chamber generates toroidal vortex in piston pocket, mixing fuel and air. 005BTC: Fuel and air is constrained to, and homogenously mixed in, central chamber. 004BTC: Spark ignites fuel and air mixture, fuel-lean combustion is violent. 000TDC: Cylinder pressure increases, combusting gas expands into backfill passage. 001ATC: Laminar flow within backfill passage returns pure air to crevice chamber. 004ATC: Fuel-lean combustion is violent and complete. Reaction extinguishes. 005ATC: Peak cylinder temperature and pressure begins dropping adiabatically. 029ATC: Combusted gasses are adiabatically cooling in combustion chamber. 030ATC: Combustion chamber first exposes thermally conductive cylinder bore. 040ATC: Central and crevice chambers transform back into single combustion chamber. 120ATC: Exhaust valve opens (actual timing depends on needs of compression stroke). 120ATC: Intake port becomes visible to combustion chamber, injects fresh filtered air. 180BDC: Intake port fully visible to combustion chamber, continues injecting fresh air.
As with the 4-stroke IPC engine, the 2-stroke IPC engine must selectively deactivate individual cylinders from service to operate at lower horsepower levels. To effectively deactivate individual cylinders in the 2-stroke IPC engine, the pressurized fresh air at the cylinder’s intake duct must be blocked at the intake plenum, and a filtered-air vacuum line of 0.25 ATM pressure must then tap into this blocked intake duct to minimize port pumping energy losses. The crankcase can be a direct source of this vacuum, though crankcase vacuum is generally saturated with oil mist. A separate vacuum canister connected to the same vacuum pump that evacuates the crankcase may provide an effective oil-free vacuum source. Similarly, the exhaust duct must be blocked at the exhaust plenum to reduce port pumping losses, and the same filtered-air vacuum canister source must then tap into the blocked exhaust duct. It is most effective if the deactivated cylinders shift EVO timing from 120ATC to 070ATC, allowing the cylinder to always remain at or above crankcase pressure, thus preventing oil seepage past the piston rings which could contaminate the combustion chamber. A positive pressure will build in the deactivated combustion chamber near TDC while not permitting significant air mass be compressed so it does not generate significant toroidal vortex friction energy losses (TKE losses), and does not generate sufficient cylinder pressure to notably increase sliding friction of the gas ported compression ring.
Full engine cycle of a deactivated cylinder in the 2-stroke IPC engine:
120BTC: Intake ports in cylinder bore become blocked by piston. 119BTC: Combustion chamber contains fresh filtered air at 0.25 ATM pressure. 070BTC: Exhaust valve closes, compression of 0.25 ATM fresh air begins. 069BTC: Fresh air begins adiabatically heating. 040BTC: Combustion chamber divides into central chamber and crevice chamber. 034BTC: Crevice chamber pumps reduced fresh air mass toward piston pocket. 030BTC: Combustion chamber becomes predominantly thermally insulating. 025BTC: Crevice chamber generates reduced toroidal vortex of air in piston pocket. 000TDC: Cylinder pressure peaks at 11 ATM of pure air. 030ATC: Combustion chamber first exposes thermally conductive cylinder bore. 040ATC: Central and crevice chambers transform back into single combustion chamber. 070ATC: Cylinder pressure drops to 0.25 ATM. Exhaust valve opens. 120ATC: Intake port becomes visible to combustion chamber, shares 0.25 ATM air. 180BDC: Intake port fully visible to combustion chamber, 0.25 ATM air in chamber.
Continuous crankcase pressure is targeted at 0.25 ATM, but it may be found 0.50 ATM is a more practical value. The oiling system will be designed for the 0.25 ATM specification, since it will function properly at any value above 0.25 ATM.
On further study of the mechanical complexity of changing valve timing at individual cylinders, as opposed to changing valve timing at all cylinders as a group, a cylinder deactivation algorithm that requires adjusting valve timing at individual cylinders is not being considered for the 2-stroke concept that is now being modeled in CAD. The 4-stroke IPC engine concept above has demonstrated the complexity of adjusting valve timing at individual cylinders, and this need not be revisited. The 2-stroke IPC engine concept will not resort to the valve timing algorithm for deactivated cylinders that was suggested above, but will retain the suggested vacuum at the intake and exhaust ducts of deactivated cylinders to minimize pumping energy loss. For this reason, the 2-stroke IPC engine concept will retain 0.25 ATM as the pressure applied to deactivated ducts. The 2-stroke IPC engine now being modeled will tentatively retain eight exhaust valves per cylinder, four being actuated by a first variable timing camshaft of fixed 160 degree duration, the remaining four being actuated by a second variable timing camshaft of fixed 160 degree duration. The first camshaft will be tasked with varying the compression ratio, the second with varying the expansion ratio. The two camshafts will be bulk actuated at the drive ends, their timing being independent from each other, each camshaft varying over a 25 degree range overlapping only at EVO 120ATC, with the combined capability of providing varible duration suitable for conditions as complex as flex-fuel (propane, gasoline, diesel, etc). When this simpler valvetrain construction is applied with hydraulic lifters at each valve stem and rollers at each cam lobe, the result may be a practical 2-stroke 3.2 liter IPC engine which generates 75 HP at 4000 RPM.
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قديم 27-01-2011, 04:29 PM
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