Fig10 - The Insulated Pulse Engine concept
of 2010 (on left) is a versatile construction intended to function throughout a wide range of operating environments.
Revision 2011 (on right) is a simpler, cost-reduced engine concept intended for use when the operating environment is more specifically defined.
[enlarge.jpg].
The Insulated Pulse Engine: A cold adiabatic engine concept By Dave
Schouweiler, updated 15May2012 This personal study describes a thermally efficient concept for
combusting fuel in an internal combustion engine. It explores adiabatic “ceramic” 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 combusted gasses into an exhaust duct for the post-processing of energy. Instead, this "cold adiabatic
engine” concept applies the principle attribute of the adiabatic engine (thermal insulation), along with principle attributes
of the Diesel engine (unthrottled induction and high compression ratio), the HCCI engine (isochoric heat addition), and the
Atkinson engine (isobaric heat rejection), such that combusted gasses are adiabatically cooled before exiting the combustion
chamber.
This is a conceptual paper, not a technical paper, as it contains
intuitive approximations and primitive constructions which require refinement. I’m not an engine designer and I don't
claim this concept works. I’ve been pondering this idea for a while now and have been wishing car manufacturers would
someday build a vehicle which contains an engine like this. Since there are no signs it will happen, and unable to find
a technical forum that will sustain a discussion on this topic, the time came to model the idea up on a computer and evolve
some answers. The computer model presented here provides the basis for creating the energy equations needed
to eventually prove or disprove this concept on paper. If the computational results look promising on paper, advanced analysis
using industry-recognized engine simulation software can next be sought. I will keep this webpage updated with my latest
findings. 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.  Fig11 –
The Insulated Pulse Engine is an evolving concept modeled as a 3.2 liter inline 4-cylinder with a bore of 100mm and a
stroke of 100mm. The 2011 revision (on left) and the 2010 revision (on right) are 2-stroke engines, each producing 65 horsepower.
The 2009 revision (not shown) is a 4-stroke producing 50 horsepower. The 2011 revision measures 270mm by 800mm
by 720mm tall. The 2009 and 2010 revisions share a platform and measure 430mm by 780mm by 730mm tall. The Insulated
Pulse Engine runs cool without a cooling system, requires no catalytic converter to operate at low pollution levels, no muffler
to exhaust quietly, and can use plastic piping to duct exhaust gasses. [enlarge.jpg].
A Brief Introduction to the Insulated Pulse Engine
The "insulated pulse-combustion engine",
abbreviated "insulated pulse engine" or "IPC engine", is an engine concept which studies the four
sources of heat export from internal combustion engines in an effort to improve fuel economy. This concept applies rapid
combustion, thermal insulation, and an extended expansion cycle as the unconventional means to achieve this goal. This engine
also applies a unique stratified combustion chamber to minimize the creation of pollution emissions, since conventional emissions
controls are not effective at scrubbing pollutants from the resulting cool exhaust gasses.
The IPC engine’s thermodynamic sequence applies rather pure forms of: 1) isentropic
(adiabatic) compression, 2) isochoric (pulsed) heat addition, 3) isentropic (adiabatic) expansion, and 4) isobaric heat rejection.
This sequence provides opportunity for comparatively high thermal efficiency, however it also brings with it the
penalty of a comparatively low average cylinder pressure (technically dimensioned as low “indicated mean effective pressure”
or low IMEP) which, when accounting for friction, delivers comparatively low volumetric efficiency (dimensioned
as low “brake mean effective pressure” or low BMEP).
The P-V diagram of an ideal Otto
cycle (not shown, but readily searchable) indicates the early Otto engine had a purely isochoric heat addition process, however
it is well recognized the modern Otto engine (modern gasoline engine) has evolved to incorporate a two-stage heat addition
process which starts out isochoric and transitions to isobaric, with the isobaric segment significantly increasing volumetric
efficiency while significantly increasing energy loss to the exhaust stream. Similarly, the P-V diagram of an ideal
(early) Diesel engine presents a purely isobaric heat addition process, but, like the Otto engine, the Diesel engine
has also evolved toward a two-stage (sometimes called "dual-cycle") heat addition process which starts out isochoric
and transitions to isobaric to improve volumetric efficiency.
The
IPC engine concept principally differs from a modern gasoline engine in that it: 1. Optimizes isentropic compression through the use of unthrottled induction (much the way a Diesel
engine does),
2. Eliminates the isobaric component of heat addition
(much the way HCCI engine prototypes do),
3. Extends isentropic
expansion beyond convention, enabling isobaric heat rejection (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 is selectively insulated using economical Fe60Ni40 alloy steel inserts to minimize heat rejection to a cooling
system. Combustion initiates, and is consumed rapidly, near top dead center (TDC), assuring the entire fuel budget performs
work on the piston through the full expansion cycle. The expansion cycle is extended beyond convention to extract additional
energy from the pressurized gasses, further reducing average combustion chamber temperature to minimize stress on the thermal
insulators, eliminating the need for a cooling system and resulting in an exhaust stream 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 turbulent
region of the combustion chamber specifically shaped (only at TDC) to support efficient, clean combustion. Fuel stratification,
in conjunction with spark ignition (or other precision ignition method), permits throttling a homogenous fuel-air equivalence
ratio within the highly reactive range of 0.40-0.80 to assure a rapid, complete combustion reaction with a practical
torque band. A fuel-air equivalence ratio below 1.00 represents the deviation of a stoichiometric ratio toward fuel-lean.
Due to stratification
constraints, crankshaft RPM is limited 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. Since select portions of the combustion chamber are thermally insulated, combustion
chamber surfaces warm up almost instantly at engine start-up, minimizing cold-start forms of pollution emissions. Gasoline,
diesel, propane, ethanol, ammonia, or most any conventional motor fuel is applicable to the IPC engine, though some will perform
better than others. When compared with naturally-aspirated 4-stroke Otto and
Diesel engines at full throttle, a similarly displaced 2-stroke IPC engine at full throttle consumes roughly a twelfth 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 each combustion event that Otto or Diesel engines of similar displacement consume at full throttle, and only one
third of the piston stroke of the 2-stroke IPC engine is used during the compression cycle. The 2-stroke IPC engine is expected
to average twice the fuel economy of Otto and Diesel engines, and will therefore generate roughly a third of the horsepower
of similarly displaced 4-stroke Otto and Diesel engines at full throttle and similar RPM. The 4-stroke IPC engine is also expected to average twice the fuel economy, but will consume roughly an eighth
of the fuel each combustion event, as half of the piston stroke is used during the compression cycle.
The cylinder displacement requirements of 2 and 4-stroke IPC engines are respectively three and four times
that of 4-stroke Otto and Diesel engines at equivalent horsepower and RPM, but the cost, weight, and space requirements of
IPC engine assemblies remain roughly comparable to Otto and Diesel engines 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 must be constructed using methods which emphasize reduction of mechanical
friction, windage friction, and port pumping losses, friction generated within the IPC engine is roughly comparable to friction
generated within equivalently powered Otto and Diesel engines. For comparative purposes, this
paper references only naturally-aspirated engines, not supercharged engines.  Fig12 – Cutaway image of the 2-stroke Insulated Pulse Engine concept of 2011.
[enlarge.jpg], [animate.gif], [video.mpg], If the "video.mpg" link is selected, a low-res sample video is quickly presented in
a new window, along with an option to download the full-resolution video file.
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 volumetric efficiency improvements (i.e.: increased peak horsepower per unit volume
of cylinder displacement), not thermal efficiency improvements. Fuel mileage gains have come from a deliberate matching of
small displacement engine to large vehicle such that an engine is simply 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
at low to moderate loads, leading to better fuel mileage than achievable with a 160 horsepower 4 liter engine in the same
large vehicle at low to moderate loads. Fuel mileage efficiencies of small displacement engines disappear when operated at high loads.
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 as small engines more routinely populate large vehicles. Atkinson engines, which are found
in some of today’s most fuel efficient cars, achieve improved thermal efficiency through an expansion process which
has reduced volumetric efficiency and which expels less heat to the exhaust duct than equivalently powered Otto or Diesel
engines. HCCI engine development programs, now popular in laboratories around the world, seek high thermal efficiency through
a combustion process which has reduced volumetric efficiency and which also expels less heat to the exhaust duct. These
two engines suggest low volumetric efficiency may provide a pathway toward significant improvement in engine thermal efficiency
and fuel mileage.
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, and to minimize mechanical shockloading of the brittle ceramic, these adiabatic engines were designed to combust
with a conventional low heat release rate. This low heat release rate promoted combustion well into the expansion cycle.
The portion of fuel which combusted later in the expansion cycle expanded at a lower compression ratio than the fuel which
combusted near TDC. The latter combusting fuel, in combination with thermal insulation, resulted in a superheating
of the combustion chamber gasses before they were expelled 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 practical 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 economy in vehicles has found minimal research interest
in the industry since the conclusion of these experiments.
The ceramic adiabatic engine experiments being developed by industry throughout 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 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.
As a 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 and pressure energy to the exhaust than do equivalently
powered Otto, Diesel, and adiabatic engines.
 Fig13 – This is a cutaway image containing only the rotating, reciprocating, and counterbalance components
of the 2-stroke IPC engine of 2011, and represents essentially all moving components within the engine, outside of ordinary
fuel and oiling functions. The energy of mechanical vibration neutralized by this counterweight scheme is redirected into
productive crankshaft output. The helix angle is reversed between the front set of counterweight gears and rear set to prevent
loading the crankshaft thrust bearings. Gear tooth loading is low, permitting the use of low-cost counterweight gears. [enlarge.jpg], [animate.gif], [video.mpg].
Elements of Thermal Efficiency
Thermal efficiency in an internal combustion engine is determined by the ratio between
the rate in which fuel energy is introduced into the engine and the rate at which heat energy is kinetically
transferred to the flywheel, with the arithmetic difference between the two representing the energy lost via four unproductive
heat-exporting pathways. Energy loss in an internal combustion engine can be minimized by
optimizing the following four heat-exporting pathways: 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 minimizing energy loss 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 minimizing energy loss 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 minimizing energy
loss 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 "mechanism
efficiency" minimizes loss of combustion energy to mechanical component friction and fluid pumping within the engine.
Fuel energy lost to component friction and pumping, dimensioned in terms of pressure loss, is called “friction mean
effective pressure” or FMEP, and represents the arithmetical difference between IMEP and BMEP.
Otto and Diesel engines introduce heat energy into the engine at a high rate
and transfer this energy to the crankshaft and to the four heat exporting pathways at a high rate. The IPC engine concept
introduces heat energy into the engine at a low rate and transfers this energy to the crankshaft and to the four heat exporting
pathways at a low rate, with the singular defining goal of the IPC engine being that the percentage of heat energy lost to
the four heat-exporting pathways is substantially lower than the percentage lost in Otto and Diesel engines. An analysis
of the IPC engine concept which uses industry-recognized engine simulation software can determine what this percentage will
be. Quantification of the IPC engine concept, using industry-accepted engine simulation software, is a future goal,
to be achieved as time and budget permits. Thermal energy
loss in Otto and Diesel engines
Insulation efficiency must be 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 a cooling system. Combustion efficiency is low in Otto and Diesel engines because high volumetric
efficiency does not allow full combustion of all fuel 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 during
a particular combustion event expands at lower efficiency than the earlier combusting fuel, the latter combusting fuel contributing
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, a length optimized only for compression, 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 fuel economy.
Thermal energy loss in 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. 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 while minimizing mechanical shockload to the ceramic, with the latter burning fuel delivering significant
heat energy loss to the exhaust duct. Only the small portion of fuel burning near TDC combusted at high 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 was found unsuitable for commercial
application.
Thermal energy loss in 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, allowing all fuel to perform work on the piston from the start of the expansion cycle to
the end. Expansion efficiency is low because useable pressure remains in the combustion chamber when the exhaust cycle begins,
resulting in the loss of 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 reliable operating
temperatures.
Thermal energy loss in 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, optimizing extraction
of pressure energy from the combustion chamber.
Thermal energy
loss in the Insulated Pulse conceptual engine
Insulation efficiency, combustion efficiency, and
expansion efficiency are all high in the IPC engine, and the constructions described below are expected to provide a notable
increase in fuel economy over adiabatic, HCCI, and Atkinson engines while combusting cleanly, without need for pollution controls.
The fourth listed efficiency, "mechanism efficiency", is conventionally a lesser consideration and has not been
factored into the equation before now. Mechanism efficiency plays a most important role in a low BMEP application like the
IPC engine, and may determine whether this engine concept can provide improved fuel economy over commercially successful engines
or not.
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 crevices such 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 must be 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. Specifically,
a unique stratified charge combustion chamber, described below, is incorporated to minimize creation of exhaust pollutants.
Combustion in the IPC engine is sufficiently unique that some form of emissions control will likely be required,
but emissions levels should be sufficiently low that incorporation
of any needed controls will not significantly affect cost or thermal efficiency.
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, high compression ratio, and the following three unconventional
functions, to achieve high thermal efficiency: Unconventional
Function #1 - Rapid "pulse" combustion
(like an HCCI engine). Unconventional
Function #2 - Thermally insulated combustion chamber (like an
adiabatic engine). Unconventional Function #3 -
Extended expansion cycle (like an Atkinson engine).
The resulting engine requires neither a cooling system
to run cool, nor a muffler to function quietly, and exhaust gasses are sufficiently cool and pressureless that exhaust ducting
can be made of plastic. Most pollution control devices do not function efficiently in the cool exhaust stream of the
IPC engine, so special effort must be made to prevent the creation of pollution emissions during combustion.
Unconventional
Function #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 the
entire fuel budget combusts at TDC and presses upon the piston through the entire expansion cycle, greatly reducing the percentage
of heat and pressure energy lost out the exhaust and lowering the average temperature of the combustion chamber. The ordinary
methods selected to achieve a high heat release rate are: 1. High compression ratio 2. Combustion chamber shaped to fully support efficient combustion 3.
Fuel-lean equivalence ratio optimized for rapid, complete reaction 4. Fuel-air charge turbulently mixed prior to ignition 5.
Combustion chamber turbulence present at time of ignition 6. Spark ignition precisely controls the combustion
envelope 7. Additional combustion chamber turbulence generated by combustion assists complete reaction 8. Thermally insulating combustion
chamber reduces quenching of reaction
While the rate of pressure rise
(dP/dt) in the IPC engine’s combustion chamber is unconventionally high (>30 bar rise per crank angle
degree vs. <10 bar/CAD in an Otto or Diesel engine), the IPC engine does not generate unusually high 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 by reducing the quantity of fuel present in the combustion chamber,
since the IPC engine must combust cleanly without the benefit of conventional pollution controls. The rapid rate of pressure rise in the combustion chamber of the IPC engine will form a supersonic combustion
wavefront which generates significantly more shockwave noise energy than the subsonic combustion wavefront in
an Otto or Diesel engine. It is expected the pulse combustion event of the IPC engine will be predictable and manageable,
and that it will generate significantly less structural excitation noise than the less predictable detonation reaction in
an HCCI engine (>50 bar/CAD). Detonation reaction noise is a recognized issue in HCCI engines. The
2011 revision of the IPC engine may be additionally susceptible to noise generation through structural excitation because
the combustion chamber contains two moving components: a piston and a reciprocating cylinder. This merits study,
but may not be a problem. The 2011 IPC engine concept possesses four elements which are expected to reduce noise susceptibility
over known HCCI constructions: 1) Combustion is timed to occur with precision near TDC minimizing piston skirt loading
and slap, 2) combustion near TDC will localize the combustion reaction to the center of the piston face which will then throttle
acoustic energy as the wavefront propagates through a restrictive perimeter region and then contacts only a
small segment of the cylinder bore, 3) the IPC engine's combustion event at full throttle consumes only one third the fuel
energy of an HCCI combustion event because only one third of the piston stroke is applied during compression, and 4) the supersonic
combustion reaction of the IPC engine has a defined propagation wavefront (initially reactive, then transitioning to an inert
wavefront) which produces less reaction chaos when compared with the unpredictable timing of the
waveless detonation reaction in an HCCI engine. An alternate consideration may find the reciprocating cylinder
of the 2011 revision provides acoustic isolation from the cylinder block, possibly reducing noise energy transmission rather
than amplifying it. If the reciprocating cylinder of the 2011 revision adds an undesirable acoustic component to
the noise equation, the fixed-cylinder IPC engine constructions of 2009 and 2010 eliminate the reciprocating cylinder
variable.
Engine misfire may occasionally cause an anomalous
stoichiometric fuel-air mixture to combust at excessive detonation pressures in the chamber. The IPC engine, like conventional engines, is constructed to
occasionally handle this type of misfire condition without damage.
A Caveat to the Compression
Ratio
Up to this point,
the IPC engine concept has been presented in an ideal form. It was mentioned above that "mechanism efficiency" is
one of the four sources of heat export from internal combustion engines. Due to low volumetric efficiency, the IPC engine
must focus on designing to maximize mechanism efficiency by minimizing both operating friction and fluid pumping inefficiencies.
Mechanism efficiency is expected to be the major component of heat export from an ideal form of the IPC engine, and if volumetric
efficiency is not enhanced through compromise of this ideal engine cycle it is possible that the IPC engine will be less
fuel efficient than existing commercial engine choices. An effort to manage volumetric efficiency will
now be explored.
The first of eight parameters listed immediately above to
promote supersonic pulse combustion is "high compression ratio". Compression ratio is also listed further
above as the first of three parameters which define high "combustion efficiency". This "first parameter"
is targeted for deliberate compromise in the IPC engine for the purpose of biasing toward improved volumetric efficiency.
Assuming fuel chemistry is a constant, fuel detonation defines a first-order upper bound for the compression
ratio while ignition flammability defines a first-order lower bound. Within the upper and lower bounds exists an
operating range of the IPC engine which is also susceptible to other constraints. The ambient operating environment may next be considered. At high and low ambient temperature limits exist thermal
expansion constraints for engine components, with a deck clearance specification between piston and cylinder head
restricting the upper bound of the compression ratio to a second-order value. At high and low barometric limits exist
cylinder pressure constraints which may restrict ignition flammability at the lower bound to a second-order value. When the compression ratio reaches the second-order upper bound, the combustion chamber volume
at TDC shrinks sufficiently to limit the volume of fuel which can be injected before NOx pollution limits are reached, ultimately
keeping the average temperature of the combustion chamber through a full engine cycle comparatively cool. When the compression
ratio reaches the second-order lower bound, the combustion chamber volume at TDC is larger, permitting a greater quantity
of fuel into the combustion chamber before NOx pollution limits are reached and thus increasing the average temperature of
the combustion chamber through an entire engine cycle, potentially reaching the thermal fatigue limit of the combustion
chamber insulators. Avoiding thermal fatigue defines a third-order lower bound to the compression ratio. Other operating constraints must also be considered. Once all constraints are considered, a practical (restricted)
upper and lower bound exists for the compression ratio. The upper limit of this restricted bound represents
the highest thermal efficiency and the lower limit bound defines the highest volumetric efficiency. Recognizing
the IPC engine concept is founded on low volumetric efficiency principles, biasing the compression ratio toward
the highest permissible volumetric efficiency provides the greatest opportunity to provide an engine with a commercially-competitive
volumetric efficiency. For this reason, the compression ratio will be biased toward the lower permissible bound unless
commercial demand permits otherwise.
The 'compression ratio" in an IPC engine represents a significantly different function
than it does in a gasoline or Diesel engine. The IPC engine employs the compression ratio to process the entire
combustion reaction while the latter engines employ the compression ratio only to initiate the combustion event. Of
the latter engines, a Diesel engine at idle provides an exception, as the abbreviated injection cycle of a
Diesel engine at idle initiates and concludes combustion near TDC, much like an IPC engine, resulting in remarkably high
thermal efficiency. Thermal efficiency of a Diesel engine drops off as load increases, since fuel injection occupies
an ever increasing segment of the expansion cycle as load increases. Only the fuel injected at TDC in a Diesel engine
expands at a high compression ratio. The portion of fuel injected later in the expansion cycle combusts at ever-dropping
compression ratios, with significant fractions of the fuel combusting at expansion ratios as low as 4:1 and
3:1 as engine loads increase. A Diesel engine at full load possesses an effective "composite compression
ratio" nearer 7:1 or 6:1. A gasoline engine has a lower "composite compression ratio" than a Diesel engine.
Since all fuel combusts near TDC in an IPC engine, the composite compression ratio is very close to the dynamic
compression ratio (DCR). Comparing the compression ratio efficiencies of an IPC engine to that of a gasoline
or Diesel engine requires comparing the composite compression ratios of each. For this reason, an IPC engine biased
to operate at the lower permissible compression ratio bound retains a compression ratio which compares
favorably to commercially available engines. It should be noted the IPC engine can potentially operate efficiently
for extended periods at idle (e.g. - to operate air conditioning in a parked vehicle), and is not constrained by
the "wet stacking" limits of idling diesel engines.
The compression ratio parameter will be revisited in
the section of this paper which introduces and develops the spark plug and other precision ignition constructions
applicable to the IPC engine.
Unconventional Function #2 - Thermally
Insulated
Combustion
Chamber
The IPC engine thermally insulates the combustion chamber fully
when the piston is at TDC. It partly insulates the combustion chamber as the piston drops away from TDC. Three reasons for
insulating are: 1) to increase thermal efficiency by minimizing heat energy loss to a cooling system during the hottest portion
of the compression and expansion cycles, 2) to burn cleanly at TDC by assuring critical combustion chamber surfaces quickly
flash to higher temperatures during compression and combustion to minimize the formation of CO exhaust emissions, and 3) to
bring combustion chamber surfaces up to operating temperature as quickly as possible at cold engine start-up to minimize exhaust
pollutants commonly associated with cold engine starts.
The combustion chamber is not fully insulated when the piston drops from TDC in order that lubricated cylinder bore
surfaces can quickly dissipate friction heat generated by direct contact with compression sealing rings.
The full extent of thermal insulation in the IPC engine is the piston contains a 3mm thick nickel-steel insulating
cap and the cylinder head contains a 3mm thick nickel-steel insulating dish. That’s it. One of these investment cast
insulators is pre-inserted into the die cast mold of an  aluminum piston, the other is pre-inserted into the mold of a cast aluminum cylinder head.
The preferred thermal insulating material
in the IPC engine's combustion chamber is an iron or steel alloy containing 40% nickel, 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 applicable. PSZ ceramic was not sufficiently durable
in the adiabatic engine experiments to become commercially applicable, though it performed remarkably well  considering the severity of the application. PSZ may perform reliably at the milder thermal gradients within the IPC engine,
particularly if applied without significant tensile loading, however mechanical shockloads related to the unconventionally
rapid "pulse" combustion event (>30 bar rise/CAD) may limit the applicability of ceramic within the IPC
engine until brittleness issues can be addressed. 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.
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, besides simply insulating, may additionally improve CO exhaust emissions, due to increased flash-warming
of combustion chamber surfaces during compression and combustion, however turbulence during compression and combustion is
expected to heat nickel steel combustion chamber surfaces sufficiently to minimize CO exhaust emissions. Selective application
of commercially available ceramic film coatings or catalytic coatings to the nickel steel combustion chamber may additionally
minimize CO emissions, if needed.
As indicated, there are materials which thermally insulate better than
the selected 40% nickel steel alloy, but ideal insulators are not expected to substantially improve engine thermal efficiency
in this application. The selected nickel steel will perform nearly as well as an ideal thermal insulator at high engine RPM,
and will only become significantly less thermally 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 remains significantly more thermally efficient than Otto and Diesel combustion chambers.
 Fig17 - The investment cast nickel-steel thermal
insulators shown here are used in the 2009 and 2010 IPC engines which incorporate poppet valves in the cylinder head.
The 2011 IPC engine uses no poppet valves and employs simplified versions of these castings.
Unconventional Function #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 and pressure, bringing the average combustion
chamber temperature down to the level where a cooling system is not required at all.
Otto and Diesel engines have evolved such that the compression and expansion cycles are matched in stroke length. The
compression cycle and the expansion cycle are each driven by significantly different processes and mathematical equations,
and their stroke lengths will seldom coincide if maximized fuel economy is the primary goal. The 2-stroke IPC engine's compression
cycle is roughly one third of a piston stroke and the expansion cycle is roughly two-thirds of a piston stroke. The 4-stroke IPC engine's compression cycle is roughly half of a piston stroke and the expansion cycle is
roughly a full piston stroke.
The IPC engine inducts unthrottled air, much like a Diesel engine,
such that it adiabatically pre-warms the inducted charge during compression to just below the auto-ignition temperature of
the fuel-air mixture, promoting rapid combustion when spark ignition is introduced near TDC. This puts the compression
ratio at roughly 18:1 if green-ammonia is selected as the fuel. The expansion ratio will be 36:1 in the 2-stroke
IPC engine to minimize heat energy loss to the exhaust duct, much the way an Atkinson engine minimizes exhaust energy loss.
The selection of 36:1 for the expansion ratio is based on the assumption that an arbitrary 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 / (36 ^ 1.4) = 1.0 bar equation.
 Fuel-Stratified Combustion Chamber
Two issues exist with the combustion process described above in the basic description of the IPC engine:
1) Complete
full-throttle combustion which combines a thermally efficient compression ratio with a non-stratified stoichiometric mix of
fuel and air generates destructive pressure levels if all fuel is combusted at TDC. As demonstrated in HCCI prototype
engines which use gasoline as the fuel, 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. Full-throttle equivalence ratios in this low range approach "lean
flammability limits" and combust incompletely, generating significant CO exhaust pollutants. Partial-throttle equivalence
ratios would drop below 0.15 and become too lean to combust.
2) With homogenously mixed combustion
reactions, there exist stagnant "quench" locations in the combustion chamber which don’t support efficient
combustion, yet which contain fuel and air. Examples of these locations include the tiny clearance volume between the O.D.
of the piston and I.D. of the cylinder bore above the compression sealing rings, and also at the segment of the head gasket
exposed to the combustion chamber. Significant HC pollution is created in these tiny locations of a homogenously inducted
combustion chamber, but the IPC engine is unable to use pollution controls which would otherwise scrub away this pollution.
The IPC
engine can resolve both the "lean flammability" issue and the "quench location" issue by stratifying the
combustion chamber into two separate regions just prior to direct fuel injection.
The combustion
chamber of the IPC engine is stratified 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" or "perimeter region" which contains air and actively
rejects fuel, and a "central combustion region" or "central region" which also contains only air when
the chamber forms, but which is optimized beginning at 8mm BTC to turbulently mix this air with direct-injected fuel
and combust cleanly.
An "annular transfer passage" or "transfer passage"
also forms at 12mm BTC and communicates between the two regions, transferring air toward the central region as the piston
rises above 12mm BTC and transferring fully combusted gasses to the perimeter region as the piston falls to 12mm ATC. The
transfer passage additionally acts to buffer the combustion process when the piston is within 0.5mm of TDC.
 Fig19 - Similar in shape to the 2011 combustion chamber,
the 4-stroke IPC engine of 2009 shows the combustion chamber at the transition position between stratified and unstratified,
12mm from TDC. The 4-stroke head assembly includes four small induction valves and four small exhaustion valves within
each cylinder. The valves are positioned such that fuel never contacts them, and therefore pollution emissions
cannot form within the crevices surrounding them. The 2-stroke IPC engine of 2010 employs the same head, with all
eight poppet valves used for exhaustion. The central combustion chamber is now called the central region, the crevice
chamber is now called the perimeter region, and the annular passage is now called the transfer passage.
The
stratified combustion chamber becomes optimally shaped for clean, fast combustion only when the piston is within 0.5mm of
TDC. A precisely timed and located source of ignition, such as spark ignition readily provides, is required to assure
combustion initiates and concludes precisely within this positional constraint.
As the combusting reaction heats
up within 0.5mm of TDC, the gasses expand beyond the central region. The combusting gasses efficiently spill into the thermally
insulated transfer passage, which fully supports combustion just like the central region, while pure air already residing
within the transfer passage is pushed, in laminar fashion, toward the perimeter region which does not support efficient combustion.
Only when the piston falls to 0.5mm after TDC do expanding combusted gasses reach the perimeter region. By this time the combustion
reaction has concluded and there is no concern for pollution development.
The perimeter region actively
keeps fuel away from combustion chamber features which do not efficiently support combustion. The volume of the perimeter
region approaches zero at TDC, whereas the volume of the central region approaches a finite value at TDC, creating an effective
air pump directed from the perimeter region toward the central region during the last 12mm before TDC. The perimeter region
actively pumps this air toward the central 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 region. The air pumping action from the perimeter region actively
constrains all direct injected fuel to the central 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 the IPC engine’s combustion chamber if it was not stratified.
 Fig20 - This image of the 4-stroke IPC engine of
2009 at TDC shows an obsolete overpressure-bypass valve intended to protect brittle ceramic insulators from stoichiometric
misfire conditions. The induction and exhaustion ducts are designed to emphasize low flow resistance rather than tuned
flow. The central combustion chamber is now called the central region, the crevice chamber is now called the perimeter
region, and the backfill passage is now called the transfer passage.
The
central region is shaped to fully support combustion, in that the surface area of the central region is comparatively low
to minimize quenching of the combustion reaction. The thermally insulated chamber surface heats up quickly during compression
and combustion to assure fuel in close proximity to the insulated material combusts fully. The central region is shaped to
generate within itself a toroidal vortex as air is pumped in from the perimeter region, assuring all fuel is in motion to
uniformly combust, the turbulence minimizing both hot and cold spots in the central region, minimizing pre-ignition and pollution
issues.
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. The
reaction rate defines an engine RPM maximum which, if exceeded, will result in incomplete combustion and pollution emissions.
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 initially combust near TDC. If pollutant generation
is to be low, quench features, such as spark plug insulation recesses, are not permissible in the central region or transfer
passage. The IPC engine operates with greatest thermal efficiency at or just below this RPM maximum, and it retains practical
levels of thermal efficiency at significantly lower RPM. A maximum RPM value of 4000 has arbitrarily been assigned to the
IPC engine for investigative purposes.
Application of the stratified combustion chamber allows the use
of gas-ported piston rings to reduce sliding friction during the low-pressure segment of the engine cycle. Since fuel is not
permitted to enter the region of the IPC engine’s combustion chamber occupied by the gas ports, HC pollution emissions
cannot form within them, and fuel cannot clog them.  Fig21
- The 2-stroke Insulated Pulse Engine of 2011. [enlarge.jpg].
Construction summary of the 2-stroke Insulated
Pulse Engine of 2011
As mentioned
above, the Insulated Pulse engine concept is developed around a 3.2 liter inline 4-cylinder reciprocating piston standard,
with nominal 100mm bore and 100mm stroke, the standard permitting comparison between evolving revisions. It should be
noted that deviating from the 100mm "square" aspect ratio may result in an improvement in fuel efficiency,
but for research and comparison, the bore and stroke specification is fixed. The actual reciprocating construction,
whether inline-4, radial-5, boxer-6, domino-8, V-10, W-12, 2-stroke, or 4-stroke, is flexible.
The IPC engine revision of 2011 is a 2-stroke reciprocating engine with a single crankshaft and 90-degree
firing interval. The 2011 engine additionally encloses each piston within a ported reciprocating cylinder, the combination
of which join to the crankshaft using separate bearing journals. Together the piston and reciprocating cylinder perform
the primary induction and exhaustion functions for the engine. The piston/cylinder assembly is installed into a conventional
cylinder block containing ported fixed cylinders which complete the induction and exhaustion function. The rotary shutter
valve (rotating drum) of 2010 IPC engine is essentially retasked as a linear shutter valve (a reciprocating cylinder) in the
2011 IPC engine. The crankcase is kept at a controlled vacuum to reduce windage energy losses and to promote oil return
from the cylinder head when the engine is turned off.
Piston Assembly
The piston assembly of the 2011 IPC engine is comprised of a thermally insulated piston, a connecting rod, and conventionally
associated components
Piston is selectively insulated to retain combustion heat within the combustion chamber while allowing the small amount
of heat which escapes past the insulator to quickly dissipate throughout the engine, where induction air flow and cool exhaust
gasses both draw this heat from the engine.
The insulated piston contains a set of compression sealing and oil control rings near the compression end, and adds
a second set of sealing and oil control rings near the crankcase end to manage crankcase vacuum. Lubricating oil is
metered from the crankcase ring set to the compression ring set via positional overlap within the bore during the course of
a full engine cycle. Ring oiling can be supplemented via metered passages which draw from the pressure-fed connecting
rod. Oiling requirements for the compression ring set is reduced from convention because the cylinder wall never contacts
combustion flame. Lubricating oil must be pressure fed through the connecting rod to the piston’s wrist
pin, since crankcase vacuum reduces oil mist within the crankcase and since the piston skirt is shrouded from direct
crankcase oil splash by the reciprocating cylinder. The crankcase is held at a vacuum for several reasons, one of the
reasons is because the piston is positioned within a reciprocating cylinder which restricts air flow below the piston, and
a vacuum reduces energy lost to air pumping through the base of the reciprocating cylinder caused by piston motion.
The sealing rings on the compression end travel across a band of twelve induction ports in the reciprocating cylinder.
To prevent wear caused by introduction of the ends of the rings to the unsupported space of a port window, the rings may be
pinned in the piston grooves, allowing them to float in position within the groove without being allowed to rotate in the
bore, keeping ring ends away from port windows.
The piston may be gas ported, since fuel
does not approach the perimeter segment of the combustion chamber occupied by sealing rings or gas ports, and therefore pollution
emissions cannot form within the gas ports, permitting low-tension sealing rings which reduce sliding friction when cylinder
pressure is low. Since the crankcase sealing ring set at the lower end of the piston never experiences pressure
above ambient, piston ring blowby from the compression sealing rings will not affect crankcase vacuum, and
since piston ring blowby in the IPC engine contains only compressed ambient air, normal levels of blowby will not affect
exhaust emissions.
Reciprocating cylinder assembly
The reciprocating cylinder assembly of the IPC engine is comprised of a reciprocating cylinder, two connecting rods, plus
components conventionally associated with a piston assembly. The cylinder assembly is connected to the crankshaft
through two connecting rods, and the crankshaft journals are located such that the stroke and phase angle of the cylinder
is not matched with piston motion. Present parameters generate a 60mm cylinder stroke with the journal’s phase
angle retarded 35 crankshaft degrees from the piston. The cylinder OD is 114mm, the ID is 100mm.
The reciprocating cylinder, alternately called a "sleeve
valve", contains two circumferential bands of twelve ports, the band nearer the compression end provides
exhaustion, the band nearer the crankcase end provides induction. The ports are shown obround to keep ring wear
low, but can take on a slight barber pole slant to further reduce ring wear.
It
is preferable that the reciprocating cylinder bore surface be comprised of a thermally-insulating material, however the level
of insulation must not permit excessive heat to build up to the point of burning the sealing rings of the piston which slide
within the cylinder. For this reason, a cast iron with compacted graphite is the preferred cylinder material, as it
has a proven lifespan, is low in cost, high in lubricity, and absorbs less heat than would a hypereutectic aluminum reciprocating
cylinder. Since a cast iron reciprocating cylinder will be heavy, a composite cylinder containing cast iron sleeve surrounded
by a structural aluminum sheath and aluminum base is a valid consideration. An aluminum sheath provides a second benefit,
in that it quickly carries heat away from the narrow band of cast iron which absorbs compression and combustion heat.
Absorbed heat is carried along the sheath where it is then dissipated into the cylinder block at the bottom and
into the cylinder head at the top.
A complication to the sleeve/sheath construction
in a 2-stroke configuration relates to the ports in the reciprocating cylinder. Since aluminum and iron have different
thermal expansion coefficients, it is not permissible to construct the cylinder with both materials at the two bands of reciprocating
cylinder ports, since the cylinder lacks hoop-strength at these locations and will deform as the materials separate over time.
For this reason, the two bands of ports will be comprised entirely of one material, and cast iron is selected.
The reciprocating cylinder will be constrained at the crankcase end at the OD, and at the compression end at the ID.
Except for the crankcase end OD measuring 114mm to provide positional constraint within the cylinder block, the reciprocating
cylinder will be stepped down to 113mm at the OD to provide a small but necessary clearance between the reciprocating cylinder
and the fixed bore of the cylinder block. This clearance will be large enough to prevent scuffing of the reciprocating
cylinder against the cylinder block, and small enough to prevent significant port leakage, based on the unique requirements
of this design.
The reciprocating cylinder assembly is installed into the cylinder block
from a conventional deck surface, much like a piston assembly is installed. The reciprocating cylinder assembly may
contain, but need not contain, the piston assembly at time of installation into the block. The reciprocating cylinder
contains sealing/oil control rings at the crankcase end to manage crankcase vacuum while assuring the reciprocating
cylinder OD will have a lubricating film of oil near the crankcase end to prevent bore wear of the cylinder block. The
sealing/oil control rings at the crankcase end can be calibrated for low friction through the full engine cycle.
It should be noted the reciprocating cylinder additionally slides against a set of sealing rings contained by the
piston, and against another set of sealing rings contained by the head. Both of these latter ring sets seal against
high combustion pressures, and therefore will generate notable friction when sealing combustion pressure. The reciprocating
piston has a 100mm stroke, the reciprocating cylinder a 60mm stroke which is phase-angle shifted from the piston. The
result is the piston rings actually slide only 60mm within the reciprocating cylinder, and the head rings slide the expected
60mm within the reciprocating cylinder, generating 120mm of total compression ring travel per half engine cycle, compared
to 100mm in a conventional engine. Friction introduced by the extra 20 percent extra ring travel distance in the IPC
engine is compensated for by using gas ported compression sealing rings, which can reduce sliding friction significantly compared
to a conventional engine, since the IPC engine contains elevated combustion chamber pressures for a briefer segment of the
full engine cycle than does an Otto or Diesel engine.
Crankshaft assembly
The crankshaft assembly consists of a nodular iron crankshaft containing five main bearings and a counterbalance assembly.
The central main bearing also comprises a thrust bearing function. Each cylinder position on the crankshaft contains
three bearing journals, a central journal to attach the piston assembly flanked by a pair of journals to attach the reciprocating
cylinder assembly. The crankshaft is fully drilled for pressure-oiling twelve connecting rods.
Due to the 90-degree firing sequence and four
inline cylinders, balance compensation for both rotating mass and reciprocating mass is included with the crankshaft assembly.
The rotating mass is compensated by a portion of conventional counterweights at each end of the crankshaft, and reciprocating
mass is compensated by the remainder of counterweight mass combined with a concentric reverse-spinning counterweight at each
end of the crankshaft. Helical gearing is used to provide quiet drive of the reverse-spinning counterweights. The helix
angle is reversed between the front set of counterweight gears and rear set to prevent loading the crankshaft thrust bearings.
Gear tooth loading is low, permitting the use of counterweight gears with reduced fabrication costs. The crankshaft
is drilled for pressure lubrication of the reverse-spinning counterweights and for splash lubrication of the crankshaft end
seals. The piston assembly and the reciprocating cylinder assembly both
employ unconventionally high connecting rod length-to-stroke ratios, permitting a closer approximation to sinusoidal reciprocating
motion than found in many Otto and Diesel engines. A high rod ratio allows the counterweight assembly to more
effectively neutralize vibration in the engine assembly than if more conventional rod ratios were employed. Engine
vibration represents mechanical energy produced by the engine which is diverted away from productive crankshaft output.
Balancing to minimize vibration assures the mechanical energy produced by the engine is directed most effectively into
productive crankshaft output.
Second-order reciprocating vibration is not generated in this construction; however the intermediate helical gearing
for the counterweight assembly has been sized and positioned to permit 2nd-order vibration compensation, should a 180-degree
crankshaft be trial-fitted at some point. If a 180-degree crankshaft is fitted, the four intermediate helical
gears would each be counterweighted to effectively perform the function of a pair of counterbalance shafts, and again, the
vibration neutralized by the counterbalance function would be effectively redirected into productive crankshaft
output.
Since there is
significant reciprocating and rotating mass, a damper is included at the front of the crankshaft to dissipate torsional vibration
reflected off the flywheel mass at the rear of the crankshaft.
Cylinder block assembly
The cylinder block assembly is comprised of a cylinder block, a maincap block, a front panel, a rear panel, and conventional
components associated with a cylinder block assembly.
The cylinder block casting contains four fixed cylinder bores, each having two circumferential bands of twelve ports
each, and an additional band of eight ports. The band of eight ports is nearest the deck (top) surface and are
included for mechanism venting requirements which are specific to this design. The band of ports nearest the crankcase
is for induction, and the center band is for exhaustion.
Since the cylinder block is designed to contain four reciprocating cylinders within
its four fixed cylinder bores, and since these reciprocating cylinders each have sealing/oil control rings contacting the
fixed bore nearest the crankcase end, with the reciprocating cylinders only contacting the fixed bores nearest the crankcase
end, the fixed bores must be constructed to handle the associated sliding friction of the reciprocating cylinders. The
cylinder block may be constructed entirely of a hypereutectic aluminum alloy to handle this, but this method may be more costly
than inserting pre-cast cylinder liners into the block mold prior to pouring the cylinder block, the integrally-cast inserts
permitting specific wear-resistant cylinder sleeves at the necessary positions of the block, while allowing the block to be
cast of a more cheaply machinable aluminum alloy. The integrally cast cylinder liners, if found to be beneficial, may
be a hypereutectic aluminum or a cast iron.
The cylinder block integrates the intake manifold and exhaust manifold into the block casting, and the block comprises
five separate levels, the top four being enclosed: 1) The uppermost enclosed level of the block is required for incidental mechanism venting specific
to this engine design. This level also provides two entryway ports at the top which enable fresh filtered
ambient air to be drawn into the block.
2) The second enclosed level from the top acts as an untuned manifold which collects exhaust gasses ejected by
each cylinder as the ports open, and which directs the exhaust at low-restriction toward an exhaust flange exiting the block.
3) The third enclosed level from the top
acts as an untuned manifold which draws fresh filtered air entering the uppermost level of the block, air which bypasses the
exhaust level at four large corner passageways, providing low-restriction filtered ambient air for each cylinder as the induction
ports open and draw filtered air into the combustion chamber.
4) The fourth enclosed level from the top is the oil reservoir for the engine.
This level, additionally, absorbs and dissipates a portion of the small amount of heat generated by sliding contact by
each reciprocating cylinder, and draws away a portion of the small amount of heat conducted into the reciprocating cylinder
by the compression/combustion processes.
5) The fifth level from the top, or more descriptively, the bottom level, is the crankcase.
The maincap block constrains the crankshaft to the cylinder block,
directs lubricating oil, and supports crankcase vacuum.
The front panel and the rear panel provide oil and vacuum sealing for the crankshaft,
provide vacuum sealing for the crankcase, and provide a mechanical structure which supports the intermediate helical gears
used for engine balance.
Cylinder head assembly
A cylinder head for an inline 4-cylinder engine is usually a single casting for the entire engine, however the 2011
IPC engine concept has taken the unusual approach of running a separate cylinder head for each cylinder. The principle
reason for this is to permit lower-cost machining of the cylinder block, since the head for each cylinder must be precision-aligned
to each bore to properly center the reciprocating cylinder within the fixed cylinder, and it is not mechanically trivial to
get all bores in a block perfectly aligned and spaced. Attempting to manufacture a single head which precisely aligns
to all four cylinder bores, when the cylinders may not be perfectly positioned, can be revisited as a future project.
The cylinder head of the IPC engine is
an aluminum casting with datum features turned on a lathe in a low-cost manner which assures the concentricity needed to keep
the reciprocating cylinder appropriately centered within the fixed cylinder bore such that tolerances assure a specified clearance
gap is maintained between the fixed and reciprocating cylinders to prevent scuffing. Since various configurations
of the IPC engine have both high compression ratio and high connecting rod ratio, there may be stationary applications,
such as irrigation pumps or electric power generators, which benefit from casting both the head and the cylinder block
from gray iron to minimize thermal expansion differentials throughout the engine assembly.
The cylinder head contains an investment-cast thermally insulating
nickel-steel dish which is pre-installed into the mold in order that it be cast integrally to the aluminum head. The
nickel-steel insert comprises the combustion chamber surface of the cylinder head, with the goal of minimizing heat lost during
compression and combustion while minimizing pollution emissions. The aluminum body of the cylinder head provides a thermally
conductive pathway for the small amount of heat which escapes through the thermal insulator to quickly dissipate into the
engine assembly, where the dissipated heat is eventually carried away by induction air and cool exhaustion gasses.
The head is affixed to the cylinder block
using six head bolts which are positioned to maximize the support of thrust forces applied to the reciprocating cylinder by
the piston assembly.
The head contains a primary set of sealing/oil control rings positioned low on the casting to manage combustion chamber
gasses, and the head contains a secondary set of oil control rings positioned higher on the casting to supply pressurized
oil to the uppermost end of the reciprocating cylinder, to prevent scuffing as the reciprocating cylinder’s bore slides
against the head. The travel path of the secondary set of rings overlaps the travel path of the primary set of rings,
providing a controlled volume of lubricating oil to the reciprocating cylinder. The pressurized oil supply directed
to the cylinder head circuitry is regulated at the oil pump to a lower pressure than the regulated supply which feeds the
crankshaft, to prevent overfeeding of the cylinder head with oil. The cylinder head also connects to a crankcase vacuum
passage which is present to draw oil from the cylinder head and return it to the sump.
Since oil pressure drops to zero instantly as the engine stops, and since
crankcase vacuum is retained for a controlled period of time after the engine is turned off, crankcase vacuum is used
to withdraw oil remaining in cylinder head passages when the engine is turned off, to prevent migration of oil past the oil
control rings and into the combustion chamber.
Spark plug and coil The shown spark plug is an unconventional
concept, with two insulated electrodes which permit reduced electrode voltage, smaller electrical insulators, and
reversible polarities to minimize sputter erosion (lifetime electrodes), with a dual autotransformer coil integrated into
the assembly and center tapped to engine ground through high resistance to help guide the spark. This construction
represents a generic spark ignition function applicable to the IPC engine for investigative purposes and is not intended
to suggest a best-design practice. The spark plug function in the IPC engine should be constructed in
a manner which minimizes crevice-type volumes such as the ceramic recess pockets associated with ordinary heat-rated spark
plugs, since recess crevices trap fuel in a location which does not support efficient combustion. Heat-rated spark plugs assure
that carbon from combustion does not build up on the electrodes, and removing the crevices removes the carbon buildup protection.
If ethanol is selected as a preferred fuel for IPC engine research, the crevice/carbon issue is reduced compared to gasoline
or diesel, since ethanol is oxygenated and does not tend to generate PM (soot) pollution. Should gasoline be selected
as the preferred fuel, carbon build-up at the spark plugs should be a concern, however, since the IPC engine is intrinsically
fuel lean, the presence of free carbon during combustion will be short-lived, helping avoid conditions which promote
carbon buildup, whether fueled by gasoline or diesel. An additional concern is the ceramic and electrode of the spark
plug must be resistant to the stresses induced by a supersonic combustion shockwave.
The IPC engine requires a precision timed and precision positioned source of ignition to minimize creation
of pollution emissions. The combustion chamber's central region and transfer passage are each shaped to fully support
combustion only at TDC, while the perimeter region does not support efficient combustion. It is important that combustion
be precisely initiated at the centermost location of the combustion chamber, allowing the stratified combustion reaction to
expand symmetrically outward and conclude before reaching the perimeter region. A spark/ethanol compression ratio (DCR) will arbitrarily be considered to reliably function in a
range between 12:1 to 14:1, the lower ratio arbitrarily assumed to bias toward overheating (when injecting fuel at allowable
NOx limits) and the higher ratio biased toward pre-ignition. Operating near the overheating threshold of 12:1 optimizes
volumetric efficiency, providing the greatest opportunity for the IPC engine to become commercially competitive. A renewable
fuel such as green-ammonia allows a spark-ignition compression ratio from 15:1 up to 20:1, with the crevice/carbon conflict
eliminated, and with greatest volumetric efficiency at 15:1. Spark/ammonia may not overheat until the compression ratio
drops to 12:1 or 10:1, however spark is unreliable at igniting ammonia below 15:1, therefore flammability limits will drive
the lower limit of the compression ratio rather than
overheating. Alternative methods of precision ignition might include constructions which replace the spark plug with an optical
window and external laser, or which inject pilot-quantities of a secondary fuel, such as diesel, at TDC to auto-ignite
and precisely initiate the combustion reaction of the stratified primary fuel. Since the IPC engine of
2011 lacks tribologically stressed locations like camshaft lobes, the engine lubricant can be an ordinary diesel-grade
oil which may be consumed during auto-ignition such that any water which may condense into the lubricating
oil during cold weather operation will not tend to accumulate. In cases where the primary fuel does not provide
sufficient lubrication for the injector, the auto-ignited pilot fuel may operate at higher rail pressures than the primary
fuel to assure injector lubrication.
Oiling system The oil reservoir is integral to the cylinder block, just above the crankshaft.
The helical intermediate
gears associated with crankshaft counterweights are each machined to allow fitment of a small auxiliary gear (not shown) which
drives the oil pump and sump/vacuum pump systems at the front and rear of the crankcase. The pumps themselves are omitted
from the model but all oiling circuitry is present. The IPC engine contains a single oil pump, as well as a front sump/vacuum
pump and a rear sump/vacuum pump. The two sump/vacuum pumps make passive contact with a pressurized oil passage in the
cylinder block in such a way as to assure they remain functionally lubricated when the engine assembly is operating at a tilt
angle which prevents access of one of the sump/vacuum pumps to the lubricating characteristics of sump oil for an extended
period.
Atop
the engine assembly, directly centered on the air cleaner cover, is the dip stick handle poking through. Forward of
the dip stick and rearward of the dipstick are two oil filler caps, which also act as oil reservoir vents. The vents
receive crankcase air delivered to the oil reservoir by the sump/vacuum pumps and contain condensers which catch and condense
aerosol oil droplets. Two vents exist instead of one vent, assuring proper operation when the engine assembly is significantly
tilted and one vent becomes occluded with reservoir oil.
Air cleaner assembly
The air cleaner assembly sits atop the engine block, providing filtered
ambient air at low restriction for induction. The air cleaner lid is held in place by three plastic nuts which
occupy a perimeter region surrounding the central dip stick and two oil filler/vent caps. The three nuts do not interfere
with operation of the dip stick and filler caps, but rather attach and thread outboard of three supports attached to the top
of the engine block, while the dipstick and fillers attach and thread inboard on the same three supports and operate independently
of the three air cleaner nuts. Engine starting and battery charging There is no electric starter
motor or alternator included with this engine assembly. Accessory mount and motor mount attachment points have been placed
on the cylinder block to apply these functions as may be required. The present concept applies this engine assembly to
a generic hybrid motor vehicle which uses a remotely located hybrid traction motor for engine starting and for battery
charging functions. Since this engine may be operated in a hybrid start-stop-restart manner, an air conditioning
pump may best be mounted to a transmission, permitting pump drive whether it is the IPC engine or the traction motor driving
the vehicle, and whether the vehicle is stationary or moving. Since this engine may alternately operate a motor vehicle
in a conventional (non-hybrid) manner, accessories and mounts may directly attach to the engine block.
Operating sequence of the 2-stroke Insulated Pulse Engine of 2011
The 2-stroke IPC engine of 2011 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
0.0mm TDC to 0.5mm ATC 4) Expansion – 0.5mm ATC to 67mm ATC 5) Induction - 67mm ATC
to 100mm BDC to 90mm BTC 6) Exhaustion - 95mm ATC to 100mm BDC to 33mm BTC
The 2-stroke
IPC engine of 2011 includes intake ports on
the lower cylinder bore and exhaust ports
on the upper cylinder bore. The operating sequence comprises: 33mm BTC: Exhaust ports
close,
compression
of fresh air and traces of exhaust begins. 32mm BTC: Fresh air begins adiabatically heating. 12mm BTC: Combustion chamber transitions to become stratified. 08mm BTC: Fuel is
direct injected toward pocket at center of piston. 07mm BTC: Perimeter region pumps fresh air toward central region, constraining
fuel. 06mm BTC: Direct fuel injection ends. 05mm BTC: Air sourced from perimeter region generates turbulence in central region. 01mm BTC: Fuel and air homogenously mixed in turbulent central region. 0.5mm BTC: Spark ignites fuel and air
mixture, combustion progresses rapidly. 0.2mm BTC: Combustion reaction expands into transfer
passage. 0.2mm ATC: Transfer
passage
forces pure air back into perimeter region. 0.5mm ATC: Combustion
reaction completes and extinguishes in perimeter region. 05mm ATC: Combusted gasses are adiabatically cooling in combustion chamber. 12mm ATC: Stratified combustion chamber transitions to become
single chamber. 33mm ATC: Conventional expansion cycle
ends, Atkinson expansion cycle begins. 66mm ATC: Atkinson cycle ends. Combustion chamber pressure reaches 1 bar. 67mm ATC: Intake ports in lower cylinder open. 68mm ATC: Vacuum forms and draws fresh air into lower third of combustion chamber. 69mm ATC: Upper 67mm of chamber contains gasses with 1/4 oxygen consumed. 95mm ATC: Exhaust
ports in upper cylinder begin to open. Intake ports are fully open. 100mm BDC: Intake ports begin to close. 99mm BTC: Lower 1/3
of combustion chamber contains air, upper 2/3 contains exhaust. 95mm BTC: Exhaust ports in upper cylinder fully open. 90mm BTC: Intake ports in lower cylinder close. 89mm BTC: Piston pushes combusted gasses in upper chamber out exhaust ports. 33mm BTC: Exhaust ports close, compression of fresh air and traces of exhaust begins.
Friction Management in the Insulated Pulse
Engine (This segment of the webpage is under construction)
Practical Application of the Insulated Pulse Engine
This webpage has presented a completed computer model and operational description of the most basic
physical form of the IPC engine concept. The IPC engine revision of 2011 is currently tasked with providing high fuel
economy only when operating from mid-throttle to full-throttle (moderate to high peak combustion pressures). The next stage of concept development (a future
stage, when time permits) will look toward configuring pencil-and-paper simulations (to be written actually in VisualBasic), using
the computer model of the 2011 IPC engine as the parametric reference, which will mathematically prove or disprove the
energy equations which comprise this engine concept when operated from mid-throttle to full-throttle.
Once the basic IPC engine concept is validated from mid-throttle to full-throttle
in the form of a mathematical proof, operation at low-throttle levels (low peak combustion pressures) will become a priority
engine design interest. It has not
yet been mentioned that combustion at low-throttle levels in the IPC engine is not a trivial matter, and is considerably more
complex than it is in Otto and Diesel engines. Because minimal fuel is consumed at low-throttle,
and because low throttle operation adds a vast array of development variables which need tending, it is reasonable to give
low-throttle operation a low priority until the basic IPC engine concept is validated. When power demand for the IPC
engine drops below mid-throttle the engine is presently tasked to shut down, allowing a hybrid power system to step in and
perform low-throttle duties. Hybrid vehicles routinely shut down and restart their gasoline engine as traffic requirements
dictate, and the hybrid vehicle's energy reservoir is ideally suited for low-throttle situations.
As described above, the IPC engine concept stratifies fuel to enable clean combustion.
The IPC engine combusts cleanly when the fuel-air equivalence ratio is nominally between 0.80 and 0.40, the actual values
determined by the selected fuel. It can be noted that a 0.80 equivalence ratio represents full-throttle and 0.40 represents
mid-throttle, and throttle position is independent of crankshaft RPM. When the equivalence ratio rises significantly
above 0.80 the reaction begins to slow and pollutants of the type which suggest oxygen deprivation begin to form. When the
equivalence ratio drops below 0.40 pollutants of the type which suggest an excessively cool reaction begin to form. As the
ratio drops further flammability limits are reached and the fuel-air mixture can no longer combust.
In the 3.2 liter 2-stroke IPC engine concept, an equivalence ratio of 0.80 translates
to generating 65 horsepower at 4000 RPM (the nominal redline) and 0.40 translates to 32 horsepower at 4000 RPM. Scaling down,
and anticipating the IPC engine will perform smoothly and reliably as low as 1000 RPM, 0.80 generates 16 horsepower at 1000
RPM, and 0.40 generates 8 horsepower at 1000 RPM. The stated horsepower values represent the necessary external load the crankshaft
must work against. If the crankshaft disconnects from the workload while engine power is applied the engine would immediately
overrev and RPM-sensing circuitry would halt fuel injection. To smoothly drop below the 8 horsepower minimum workload limit
requires a low-throttle capability not intrinsic to the 2011 IPC engine concept. Since low-throttle operation is expected
to become a future consideration, it is reasonable to briefly consider some solutions.
One method for dropping power output further, without revising the IPC engine concept, involves de-selecting individual
cylinders by halting fuel injection to them. If a pair of the four cylinders are de-selected, a 4 horsepower minimum
output at 1000 RPM is achieved, however engine roughness will increase slightly. If a third cylinder is de-selected a 2 horsepower
minimum output at 1000 RPM is achieved, resulting in noticeably bumpy power application to the tires. Extrapolating such that
a single cylinder is fired "every other" crankshaft revolution, output drops to 1 horsepower at 1000 RPM, however
the flywheel would need a substantial mass or complexity if the engine is to avoid stalling at this rarified firing rate.
Also, cylinder chilling caused by the interleaved deselection of an individual cylinder might allow pollution emissions to
form, particularly in colder ambient conditions. Just as the thermally insulated combustion chamber in the IPC engine
warms up quickly, it cools down quickly as well. Due to the potential for pollution generation, until experimentation
proves otherwise it will not be permitted to fire a single cylinder "every other" crankshaft revolution in the IPC
engine. Substantial flywheel mass would be prohibitive, since the engine would perform sluggishly at higher RPM, but
a flywheel complexity increase could include a twin flywheel stack in which one flywheel is conventionally fastened to the
crankshaft flange and the second flywheel is attached to the first through a 1.6:1 gear hub which activates when the crankshaft
drops below 1600 RPM and which directly locks to the first flywheel above 1600 RPM. Should minimum power application
still prove rougher than targeted, a third flywheel could cascade onto the second flywheel.
Having established a lower limit of 2 horsepower at 1000 RPM for the initial IPC
engine concept, a look at parasitic energy losses is in order. Since the horsepower rating includes parasitic losses
when all cylinders are operating, the horsepower ratings when individual cylinders are deselected are extrapolations which
assume the parasitics scale to zero as the number of active cylinders scales to zero. This is not actually the case.
In fact, when two cylinders are deselected they assume the parasitic losses formerly handled by the deselected cylinders.
Rather than the IPC engine with two cylinders deselected generating a 4 horsepower minimum output at 1000 RPM, it might actually
generate only 3 horsepower. Scaling to three deselected cylinders, and assuming each cylinder is found to support
0.5 horsepower of parasitic losses at 1000 RPM, 2.0 - (3 * 0.5) = 0.5 horsepower could reasonably be assumed the
minimum single cylinder horsepower output. Whether an 8 horsepower minimum, or with complex flywheel a 0.5 horsepower
minimum, the lower limit of the initial IPC engine concept can be further managed to zero, and therefore to a stable no-load
idle, if the traction motor (this assumes a hybrid-electric vehicle application) consumes the surplus power in the form
of either a 6kW or 400W electrical load, driving accessories as needed or simply dissipating the surplus as
heat.
This demonstrates the IPC engine concept of
2011, if mathematically validated, can be practically applied. Having clarified this, it is time to briefly look
at the possibilities associated with further development of low-throttle functions.
A Simplified
Idle
Selecting green-ammonia as the fuel provides a low-cost, low-throttle,
low-pollution opportunity which will now be presented. This paper does not address the unique safety issues
intrinsic to each fuel, it only addresses the functions and limitations of cost-competitive
fuels when applied to the IPC engine. Green-ammonia is not flammable in open air, is found naturally
in the environment, and is manufactured by combining water, air, and electricity. This low-throttle opportunity may
extend to some carbon-based fuels, particularly oxygenated fuels, if experimentation demonstrates PM (soot) pollutants
are not generated. This low-throttle "green-ammonia" algorithm
requires that each cylinder is allowed to operate within two load ranges: A first load range operates
from half-throttle to full-throttle and applies the stratified combustion chamber discussed throughout this paper. A
second load range operates from idle to quarter-throttle and does not apply the stratified combustion process, but instead
applies direct injection at metered peak flow rate at TDC in combination with a multispark ignition process. Due to
insufficient fuel/air mixing, this process may not permit operating a combustion chamber from quarter-throttle
to half-throttle without the potential for pollution generation (e.g., oxygen starvation within the flame kernel), however
this restriction is not an issue if individual cylinders within the engine are permitted to operate at different
load levels simultaneously. For instance: To avoid harshness in power delivery, two cylinders 180 crankshaft
degrees apart may operate at one matched load level while the remaining pair of cylinders may operate
when required at a different matched load level which is independent of the first two cylinders. This "pairing"
algorithm can provide a seamless spectrum of smooth power delivery from idle to full throttle using an ordinary flywheel
and without concern for managing the dissipation of surplus power at idle as was described above. As an application example: When accelerating a vehicle from a
stop, all four cylinders will operate equally from idle to quarter-throttle, at which point two cylinders will bump-up
instantly to mid-throttle (moderate peak combustion pressure) while the remaining two cylinders jump instantly to idle
(low peak combustion pressure). The two mid-power cylinders will hold steady while the two idling cylinders climb
to quarter power, at which point they hold steady and the mid-throttle cylinders then climb to 3/4-throttle. Once
two cylinders are at 3/4 throttle and two are at quarter-throttle, all cylinders jump to mid-throttle and rise to
full throttle (high peak combustion pressure) at the same rate. Note that throttle position in the IPC engine does not
indicate horsepower, it indicates torque, making this scheme applicable through the full power band of the engine,
providing versatility during motor vehicle operation, whether hybrid or conventional.
A Low-Throttle
Alternative It may be discovered that the green-ammonia fuel described immediately above requires a sufficiently
elevated compression ratio that mechanical interference between the piston and head become problematic as operating conditions
vary, or that the tiny combustion chamber volume required to sufficiently compress ammonia operates with sufficiently low
average operating temperature and low volumetric efficiency that the engine is not commercially competitive. The preferred
fuel may then shift toward propane or ethanol, or may further shift toward gasoline or diesel, permitting respectively larger
combustion chamber volumes which are associated with higher average operating temperatures and higher volumetric efficiencies.
This fuel shift may bring with it some low-throttle pollution emissions issues which will require emissions controls. As noted previously, modern automotive emissions
controls are designed to operate efficiently when exhaust temperatures are high, however the IPC engine's exhaust stream is
comparatively cool. While hydrocarbon (HC) and carbon monoxide (CO) pollutants must be prevented when operating at low-throttle
because they generally require a hot exhaust stream to be scrubbed clean, oxides of nitrogen (NOx) and particulate (PM) type
pollutants may be readily scrubbed clean in the IPC engine's cool exhaust stream. PM (soot) emissions (applicable only
to some fuels) may be resolved (if necessary) using a twin-path particulate filter equipped with a flow valve to route 99%
of the exhaust to one path for particulate collection while 1% of the IPC engine's intrinsically-oxygenated exhaust is routed
to the opposite path for electrical baking and catalytic oxydation of collected particulate. The particulate filter
valve would switch back-and-forth as required to keep exhaust backpressure low.  Fig30 - Revision 2012 of this concept exists
only to introduce low throttle configurations available to the IPC engine. It introduces an optional
idling port and associated idling plug (light yellow), the function being either active (fully retracted) or inactive
(fully inserted). Detail of an actuating linkage for the plug is omitted from this introduction. [video.mpg].
Idling
Port  The Atkinson cycle is only tuned to provide best fuel-efficiency at a single throttle
position, and in this case the chosen position is full-throttle. Operating the IPC engine below the optimal throttle position
results in a slight degradation of fuel efficiency in the form of induction port pumping losses which can be quantified
by the level of vacuum which forms in the combustion chamber prior to the start of the induction cycle. As the throttle level
slews further toward mid-throttle, quarter-throttle, and idle, the level of vacuum increases. Deviating below full-throttle
results in induction of more ambient air than is necessary, resulting in consumption of more pumping energy than is necessary. With the Atkinson cycle optimized only at full throttle, pumping energy loss at
idle can comprise a significant portion of "mechanism efficiency" at idle, however this can be downplayed somewhat
because the intrinsically low rate of fuel consumption at idle means the fuel cost of this inefficiency is quite low. At idle,
chamber vacuum begins to form near 33mm ATC and vacuum peaks near 67mm ATC, with the pumping energy inefficiently released
at the start of the induction cycle. Since there may be applications in which the engine spends an inordinate amount of time
near idle, fuel savings can quickly accumulate if this low-throttle inefficiency is
resolved.
A low-cost resolution to Atkinson inefficiency
at idle is found in the 2012 revision of the IPC engine by adding an extra port in the cylinder block's bore which deactivates
the Atkinson function when open, and activates the Atkinson function when plugged. This optional port is so-named  the "idling port". It is strategically positioned
between the induction port and the exhaustion port such that, when open, it provides an additional connection between the
combustion chamber and the exhaustion plenum for the purpose of preventing accumulation of cylinder vacuum between 33mm
ATC and 67mm ATC, effectively mirroring the compression cycle and expansion cycle to the same 33mm stroke length. An
"idling plug" can then be installed into each of the idling ports such that, when the idling port becomes blocked
by the plug, the Atkinson function is restored for optimal efficiency at full-throttle. This plug is most economically
applied in the form of a two-position plunger which is pressed flush to the cylinder block bore to apply the Atkinson segment
of the expansion cycle, or is retracted outward from the cylinder block bore to efficiently disable the Atkinson segment of
the expansion cycle. These plungers can be selectively placed radially around each cylinder block bore, or a simplified
construction may place them parallel to each other (as shown) to permit use of a simple actuation mechanism. Since idling
plungers are neither exposed to significant pressure nor are they prone to wear, and since fitment precision is comparatively
relaxed, the idling mechanism can be low in cost.
Expansion
Buffering
 Expansion buffering is an optional function which broadens the range of throttle positions in which the Atkinson cycle
operates near peak fuel efficiency in the IPC engine. Expansion buffering functions independently of the idling ports,
and if idling ports are omitted from an engine it requires only the addition of some tiny ports cut into the cylinder block
bores which connect to the exhaustion plenum, with no additional engine components. Since the 2012 revision of the IPC
engine is intended to provide only an introduction to idling constructions, and since it is expected that expansion buffering may
often be applied in conjunction with idling ports, expansion buffering will be presented here integrated with the optional idling
plunger function rather than as a stand-alone function. Adding an optional aperture to
the lowest segment of the idling plunger (refer to sculpting added to the lower set of plungers at left) allows mid-throttle
Atkinson operation to become more fuel-efficient by venting excess cylinder vacuum. This optional buffering also allows
recalibration of the Atkinson cycle to optimal levels below full-throttle, thus permitting slightly increased volumetric efficiency,
as the recalibration allows venting excess pressure (at full-throttle) to the exhaustion plenum just prior to the start of
the start of the induction cycle for the purpose of normalizing combustion chamber conditions over a range of throttle positions
such that induction pumping losses are minimized.
Carrying revision 2012 to a conclusion might include development of a fully-adjustable Atkinson function tuned to perform
optimally at all throttle positions from idle to full throttle. A fully-variable Atkinson function would provide sufficient
departure from the function to warrant another name.
Cold Weather Icing in the Insulated Pulse
Engine
Icing in the IPC engine
concept is a fun consideration which deserves a glance.
Operating while idling or with deselected cylinders in cold ambient conditions can mix
within the exhaust plenum the combination of warm humid exhaust and cold uncombusted air, potentially resulting
in frost build-up which can restrict or halt exhaust flow. Disconnecting the induction ports of deselected
cylinders from the induction plenum and connecting them to the exhaustion plenum to prevent introduction of
cold air can help prevent icing. Exhaustion plenum heaters may also be required. Due to unconventionally cool
operating temperatures, water condensation and the resulting corrosion may be a more significant problem in the IPC engine
than in conventional engines. An electric heating function within the engine block may be required during
cold weather operation to quickly bring the IPC engine's oil up to an operating temperature which prevents the accumulation
of water condensation. Alternately, lubricating oil may be retained in a small "recirculating oil" reservoir within
the engine block and may be additionally tasked with providing lubrication for the fuel injectors
if a non-lubricating fuel is employed. In conjunction with a larger "fresh oil" tank external to the
engine which keeps the recirculating reservoir topped-off, trace "metered" amounts of recirculating oil can
be injected along with fuel to minimize water accumulation in the recirculating oil, keeping recirculating
oil fresh and reducing the potential for crankcase corrosion. The external fresh oil tank would still need
to be refilled every few months, but oil would never need to be changed, only the oil filter. Since there are no tribologically
stressed lubricating surfaces in the IPC engine, low-cost diesel fuel may be an ideal lubricating oil for the IPC
engine.
Construction summary of the 2-stroke Insulated Pulse Engine of 2010
 The Insulated Pulse Engine revision 2011 (presented
throughout this webpage) has described a simple, cost-reduced engine intended for use when the operating environment is narrowly
defined (i.e., fuel parameter is fixed), and revision 2012 (briefly introduced above) provided a glimpse toward future development,
should revision 2011 prove to be a valid concept. The Insulated Pulse Engine concept of 2010 (presented next) is an earlier
construction intended to function throughout a wide range of operating environments. Particular attributes of revision 2010
include the ability to fully adjust the timing of the exhaustion cycle, allowing active adjustment of the compression ratio
(DCR) without varying combustion chamber volume. As can be expected in evolving concepts which explore unconventional processes, a number of constructions within
the 2010 IPC engine have evolved significantly since originally modeled. The twin crankshafts intended to reduce piston skirt
thrust friction evolved toward a single crankshaft in 2011, in part because it was realized the pulse combustion process does
not generate the elevated levels of piston skirt thrust found in Otto and Diesel engines. The tuned exhaust manifolds originally
presented have evolved into the currently shown "low-turbulence" manifold design, and have further evolved into
the in-block exhaustion plenum of 2011 which emphasizes only acoustic damping and high flow at lowest cost. The intake manifold
of 2010 was similarly cost-reduced in 2011, becoming an in-block induction plenum emphasizing acoustic damping
and low-restriction air filtering. From the start it was recognized that timing belts are unsuitable in engines with
a piston/poppet valve "interference" construction like the IPC engine, however the timing belts were quicker to
deploy than timing chains with internal adjusters, and both timing belts and timing chains were eliminated for 2011.
A unique form of rotary induction valving presented with this revision restricts the depth and location of head fastening
threads in the 2010 block deck, and the unique form of reciprocating induction valving in the 2011 IPC engine resolved that
limit.
The operating sequence of IPC engine revision 2010 is nominally the same as revision 2011 described above. The
principle difference is the means of porting. Revision 2010 uses a rotary shutter valve for induction rather than the reciprocating
cylinder valve in revision 2011, and revision 2010 uses poppet valves in the head for exhaustion rather than the reciprocating
cylinder valve in revision 2011. Exhaustion in revision 2010 is fully adjustable in both timing and duration, with four valves
controlling EVO and four valves controlling EVC. Exhaustion timing is used to adjust the dynamic compression ratio in the
2-stroke revision of 2010, whereas exhaustion timing is not adjustable in revision 2011. Induction is not adjustable in the
2-stroke IPC engines presented here. Details of the 2-stroke IPC engine
of 2010, published previously on this webpage, will not be reviewed at this time.
 Fig35 - Closeup cutaway image of the 2-stroke
Insulated Pulse Engine concept of 2010 showing rotary induction valves (light magenta) which operate at a close clearance
to the engine block to prevent friction and wear while permitting the descending piston to readily draw in filtered air before
rotating to prevent the return of air to the induction plenum when the piston begins to rise. [ ].
Construction summary of the 4-stroke Insulated Pulse
Engine of 2009 The 4-stroke IPC engine of 2009 provided the mechanical platform for the 2-stroke IPC engine of
2010 briefly described above. The 4-stroke cylinder block differs in that it lacks ports in the cylinder bores
and also lacks the associated rotating shutter valves of the 2-stroke. The crankshafts differed in that the 4-stroke
is a 180-degree design and the 2-stroke is a 90-degree design. The cylinder head is the same, but the camshafts
and the camshaft timing gears are calibrated differently between the 2-stroke and 4-stroke engines. In the 2-stroke
all eight poppet valves are tasked for exhaustion (4 for EVO, 4 for EVC), in the 4-stroke, four are tasked for induction and
four for exhaustion. Induction and exhaustion valve timing is independently adjustable in the 4-stroke IPC engine. Details of the 4-stroke
IPC engine of 2009, published previously on this webpage, will not be reviewed at this time.
Analysis of the Insulated Pulse Engine
Concept The Insulated
Pulse Engine revision of 2011 (presented throughout this webpage) has described a fuel-efficient engine concept. The
IPC engine revision of 2012 (briefly introduced above) provides a glimpse toward future development needs, should revision
2011 eventually prove valid. The IPC engine revisions from 2009 and 2010 remain of interest since they provide alternative
thought which may be helpful in resolving shortcomings which crop up during analysis of revision 2011. With mechanical
CAD modeling completed for the IPC engine, the next task, a future task to be commenced when scheduling permits, revolves
around calculating the operational physics of this engine concept. Specifically, the task will be to construct the energy
equations which define the steady-state operation of this engine, and then apply the equations from mid-throttle to full-throttle
using revision 2011 of the IPC engine concept as the parametric reference. This webpage will become
a technical paper rather than a conceptual paper when these energy equations are constructed, computed, and published,
and this concept will then become ready to present for advanced analysis.
While specialized software
exists for running advanced engine simulation studies, access to a license for this simulation software (AVL Fire, GT-Suite,
Ricardo Wave, etc) is expensive and training is required. As stated above, the next task is to create a no-cost
pencil-and-paper grade preliminary engine analysis written in VisualBasic, to be commenced when time permits, with
the progress posted here. Once the VisualBasic analysis is completed and presented, should the results look promising, a follow-up
study using the aforementioned industry-recognized simulation software can be sought out for the purpose of gaining a
more rigorous result. Whether analysis shows the engine concept is functional or faulty, the results will be presented and
retained on this webpage. |
