(12) INTERNATIONAL APPLICATION PUBLISHED UNDER THE PATENT COOPERATION TREATY (PCT)(19) World Intellectual Property Organization
International Bureau

11111111111111111111111111111111111111111111111111111111111111111111111111111111111111111111111(10) International Publication Number(43) International Publication Date

29 March 2012 (29.03.2012)

PCT

WO 2012/040431 A1(51) International Patent Classification:
F02B 19/00 (2006.01)-(81) Designated States (unless otherwise indicated, for every
kind of national protection available): AE, AG, AL, AM,
AO, AT, AU, AZ, BA, BB, BG, BH, BR, BW, BY, BZ,(21) International Application Number:
CA, CH, CL, CN, CO, CR, CU, CZ, DE, DK, DM, DO,
PCT/US20ll/0527ll
DZ, EC, EE, EG, ES, FI, GB, GD, GE, GH, GM, GT,
HN, HR, HU, ID, IL, IN, IS, JP, KE, KG, KM, KN, KP,(22) International Filing Date:
KR, KZ, LA, LC, LK, LR, LS, LT, LU, LY, MA, MD,
22 September20ll (22.09.2011)
ME, MG, MK, MN, MW, MX, MY, MZ, NA, NG, NI,
English(25) Filing Language:
NO, NZ, OM, PE, PG, PH, PL, PT, QA, RO, RS, RU,
RW, SC, SD, SE, SG, SK, SL, SM, ST, SV, SY, TH, TJ,
English(26) Publication Language:
TM,TN,TR,TT,TZ, VA, UG,US, UZ, VC, VN,ZA,(30) Priority Data:
ZM,ZW.
61/386,086 24 September 2010 (24.09.2010)US
61/386,605 27 September 2010 (27.09.2010)
US (84) Designated States (unless otherwise indicated, for every
kind of regional protection available): ARIPO (BW, GH,
61/436,300
26January20ll (26.01.2011)
US
GM, KE, LR, LS, MW, MZ, NA, SD, SL, SZ, TZ, UG,
61/470,566
1April20ll(Ol.04.20ll)
US
ZM, ZW), Eurasian (AM, AZ, BY, KG, KZ, MD, RU, TJ,
27June2011(27.06.20ll)
US
61/501,368
TM), European (AL, AT, BE, BG, CH, CY, CZ, DE, DK,(71) Applicant (for all designated States except US): SCUDEE, ES, FI, FR, GB, GR, HR, HU, IE, IS, IT, LT, LU,
ERI GROUP, LLC [US/US]; 1111 Elm Street, Suite 33,
LV, MC, MK, MT, NL, NO, PL, PT, RO, RS, SE, SI, SK,
West Springfield, MA 01089 (US).
SM, TR), OAPI (BF, BJ, CF, CG, CI, CM, GA, GN, GQ,
GW, ML, MR, NE, SN, TD, TG).
(72) Inventors: BRANYON, David, P.; 6220 Culebra Road,
San Antonio, TX 78238 (US). HOAG, Kevin, L.; 2029 Published:
E. Mifflin Street, Madison, WI 53704 (US). SCUDERI,
with international search report (Art. 21 (3))
Salvatore, C.; 108 Sunset Drive, Westfield, MA 01085
before the expiration of the time limit for amending the(US).
claims and to be republished in the event of receipt of(74) Agents: FILDES, Christopher, J. et al.; Fildes & Outamendments (Rule 48.2(h))
land, P.C., 20916 Mack Avenue, Suite 2, Grosse Pointe
Woods, MI 48236 (US).

---- -----------------------------------------------------------------------------------------;;;;;;;;;;;;;;;-(54) Title: TURBOCHARGED DOWNSIZED COMPRESSION CYLINDER FOR A SPLIT-CYCLE ENGINE
104

---

;;;;;;;;;;;;;;;--;;;;;;;;;;;;;;;(57) Abstract: A split-cycle engine includes an expander, the expander including an expansion piston received within an expansion cylinder. A compressor includes a compression piston received within a compression cylinder. A crossover passage interconnects the compression and expansion cylinders. An intake manifold is connected to the compression cylinder. A boosting device
providing a 1.7 bar absolute or greater boost pressure level is connected to the intake manifold. An intake valve is disposed between the intake manifold and the compression cylinder. The intake valve closing is timed to provide a compressor volumetric efficiency of 0. 75 or greater. A compressor displacement volume is sized relative to an expander displacement volume such that the
combination of compressor displacement volume and boost pressure level provides an expander volumetric efficiency relative toambient conditions that is 0.90 or greater.

wo 2012/040431

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1

TURBOCHARGED DOWNSIZED COMPRESSION CYLINDER FOR A SPLITCYCLE ENGINETECHNICAL FIELD
The
combustion

present

inventionengines.

Morerelates

specifically,tointernal

thepresent

invention relates to a turbocharged split-cycle engine that5

divides the four strokes of aconventional Otto cycle overtwointake/compression

pairedcylinders,(compressor)

and

one

one

power/exhaust

cylindercylinder(expander),connected by a crossover port, with each of the four strokes
being completed in one revolution of the crankshaft.

BACKGROUND OF THE INVENTION

10
For
engine"aspurposes

used

in

of

the

clarity,present

the

15

(or

Otto

power)

andpiston/cylinder

cycle

(the

exhaustintake,strokes)combinationof"conventional

applicationinternal combustion engine wherein all
well-known

term

four

refersthecrankshaft
20( CA) ) ,angle

(7 2 0anddegrees

two

engine.

CA)

full

are

expansioncontained

in

Each

requires one half revolution of the crankshaft
crank

to

each
stroke(180 degrees

revolutions

requiredanstrokes of thecompression,
are

to

of

the

completetheentire Otto cycle in each cylinder of a conventional engine.
Also,

for

purposes

definition is offered forofclarity,thefollowing

the term "split-cycle engine" as

may be applied to engines disclosed in the prior art and as
referred to in the present application.
25

A split-cycle engine comprises:
a crankshaft rotatable about a crankshaft axis;aslidablyreceived

operatively
30

compressor

including

withinconnected

toacompressionacompression

thecrankshaft

pistoncylinder
such

that

and
the

compression piston reciprocates through an intake stroke andwo 2012/040431

PCT/US2011/052711

2acompression

strokeduringasinglerotationofthecrankshaft;slidably5

an expander including an expansion(power)received

cylinderoperatively
expansion
andanwithinconnected

pistonexhausttoanexpansionthecrankshaft

reciprocatesstrokethroughduringasuchpiston
and

that

the

an

expansion

strokesinglerotationofthecrankshaft; anda10crossoverpassageinterconnecting

compression and expansion cylinders,
including

at

least

a

crossoverthethe crossover passage

expansion(XovrE)

valve

disposed therein, but more preferably including a crossovercompression (XovrC) valve and a crossover

expansion (XovrE)

valve defining a pressure chamber therebetween.
15

UnitedStates

2003 to Carmelo J.

patentScuderi

6,543,225

granted April(the Scuderi patent)

8,

and UnitedStates patent 6,952,923 granted October 11, 2005 to David P.
Branyon

20

et

al.

(the

extensive

discussion

engines.

In

Branyonof

addition,

patent)each

split-cycle
theScudericontains

andandsimilarBranyonantype

patents

disclose details of prior versions of engines of which thepresent invention comprises a further development.
ScuderipatentandtheBranyonpatentare

Both theincorporated

herein by reference in their entirety.
25

Referringembodiment of a

to

FIG.

1,

a

prior

art

exemplary

split-cycle engine of the type similar to

those described in the Branyon and Scuderi patents is shown

30

generally by numeral 10.

The split-cycle engine 10 replaces

two

of

adjacent

cylindersaconventional

enginewithacombination of one compression cylinder 12 and one expansion
cylinder 14.

A cylinder head 33 is typically disposed overan open end of the expansion and compression cylinders 12,
14 to cover and seal the cylinders.

wo 2012/040431

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3

The

four

strokes

of

the

Otto

cycle

are

"split"over the two cylinders 12 and 14 such that the compression
cylinder 12, together with its associated compression piston
20,
5

perform

compression

theintakecylinder

12

and
and

piston20( 12,2 0)),

referred to as the compressor
cylinder 14,

compression

strokes(thecollectively

being

and the expansion

together with its associated expansion piston

30, perform the expansion and exhaust strokes(the expansion

cylinder 14 and piston 30 collectively being referred to as
10

theexpander

(14,

30)).

The

Otto

completed in these two cylinders 12,

cycle

is

therefore

14 once per crankshaft

16 revolution (360 degrees CA) about crankshaft axis 17.
During the intake stroke, intake air is drawn into
the15

compression(port)

cylinder12

19 disposed in the
inwardthroughanintakecylinder head 33.

into

the

cylinderandAn

manifoldinwardly

opening(opening

towardthepiston)poppet intake valve 18 controls fluid communication

between the intake manifold 19 and the compression cylinder
12.
20

The intake air is approximately at atmospheric pressurein the intake manifold.
Duringpiston

25

20

the

compression

pressurizes
the

the

air

air charge

stroke,thecharge

and,

into

the

compression
upon

XovrC

opening,

drivescrossover passage(or port)

22,

head 33.

This means that the compression cylinder 12 andwhich is typically disposed in the cylinder

compression piston 20 are a source of high pressure gas tothe crossover passage 22,

which acts as the intake passage

for the expansion cylinder 14.
more
30

crossoverpassages22

In some embodiments,
interconnect

thetwo orcompression

cylinder 12 and the expansion cylinder 14.
The volumetric (or geometric) compression ratio ofthecompression cylinder 12of split-cycle engine

for split-cycle engines in general)

10

(and

is herein referred to aswo 2012/040431

PCT/US2011/052711

4

the"compression

ratio"

ofthesplit-cycle

engine.

The

volumetric (or geometric) compression ratio of the expansion
cylinder 14

of split-cycle engine

engines in general)
5(and forsplit-cycleis herein referred to as the "expansion

ratio" of the split-cycle engine.

The compression ratio ofaart

cylinder

is

well

known

enclosed (or trapped)
recesses)whenbottom dead
10

10

a

in

the

as

the

ratio

volume in the cylinder

pistoncenterthereinposition

the

to

the

(including allreciprocating

(BDC)

of

is

at

its

enclosed volume(i.e.,

clearance volume)in the cylinder when the piston is

at

top

(TDC)

its

dead center

split-cycle

enginesasposition.

defined herein,

Specifically forthevolumeofthecrossover passage(s) is not included in the determination ofthe15

compression

ratio

of

a

compression

cylinder.

Also,

specifically for split-cycle engines as defined herein,
volume of the crossover passage (s)is notdetermination

ratio

of

the

expansion

theincluded in theofanexpansion

cylinder.
Due to very high compression ratios
20

1,

30

to

opening(opening outward away from the cylinder and piston)poppetpassage

25

40

to

1,

compression
inlet

25

is

or

greater),(XovrC)

valve

used

to

an

20 to

outwardly

crossover1,(e.g.,

24

at

the

compression cylinder 12 into the crossover passage 22.

Due

1,

or

greater),

an

flow

crossoverfromto very high expansion ratios

controlthe(e.g., 20 to 1, 30 to 1, 40 to

outwardly

opening

poppetcrossover

expansion (XovrE) valve 26 at the outlet 27 of the crossover
passage 22 controls flow from the crossover passage 22 into
the expansion cylinder 14.
30

The actuation rates and phasingof the XovrC and XovrE valves 24,
thecrossover

26 are timed to maintainpressureinpassage22

atapressure(typically 20 bar absolute or higher,highminimum

e.g.,

40 to

wo 2012/040431

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550 bar, during full load operation)

during all four strokesof the Otto cycle.
Atleast onefuelinjector 2 8 injects

fuel

into

the pressurized air at the exit end of the crossover passage
5

22 in correspondence with the XovrE valve 26 opening, which
occursshortly before expansion piston 30

dead center position.
thepressure

expansion
10

inAt this time,crossovercylinderminimum pressure

14

isin thepassagehigh,reaches

its

top

the pressure ratio of
22

due

to

to

the

the

crossover passage

pressure

fact
is

that

in
the

typically 20

bar absolute or higher at full engine load and the pressure
intheexpansion

cylinderduringtheexhausttypically about one to two bar absolute.

strokeis

In other words,

when XovrE valve 26 opens, the pressure in crossover passage
15

22

is

substantially higher than the pressure

cylinder 14(typically in the order of 20 to 1 or greater atfull engine load).

This high pressure ratio causes initial

flow of the air and/or fuelcylinder
20

reach

14

the

flow.

high

speeds.

speedofsound,

The

dead

entering
25

conditions.

charge to flow into expansionatcylinder 14
topin expansion

air I fuelcharge

These

which

is

usually

high

flow

speedsreferredtoenterstheshortly after expansion piston 30

centerpositionslightly

before(TDC),
TDC

although
under

as

sonic

expansion

reaches

it
some

can

may

its

begin

operatingAs piston 30 begins its descent from its top

dead center position,

and while the XovrE valve 26 is still

open, spark plug 32, which includes a spark plug tip 39 that
protrudes into cylinder 14,

is fired to initiate combustionin the region around the spark plug tip 39.
30

Combustion canbe initiated while the expansion piston is between 1 and 30
degrees CA past its top dead center
preferably,
pistonis(TDC)

position.

More

combustion can be initiated while the expansion
between5

and 25degrees

CA past

its

top

dead

wo 2012/040431

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6

center(TDC)

position.

Most preferably,

combustion can beinitiated while the expansion piston is between 10 and 20
degreesCA

past

Additionally,
5

its

top

combustion

ignition devices

The

centerbeinitiatedmay

and/ or methods,

microwave ignition devices
methods.

dead

sonic

position.

throughothersuch as with glow plugs,

or through compression ignition

flow

particularly advantageous

(TDC)

of

the

air/fuel

chargeisto split-cycle engine 10 because

it causes a rapid combustion event, which enables the split10

cycle engine 10 to maintain high combustion pressures eventhough ignition is initiated while the expansion piston 30
is descending from its top dead center position.
The XovrE valve 26 is closed after combustion is
initiated

15

but

beforetheresultingenter the crossover passage 22.

combustion

event

can

The combustion event drives

the expansion piston 30 downward in a power stroke.
Duringtheexhauststroke,exhaust

gases

are

pumped out of the expansion cylinder 14 through exhaust port
35 disposed in cylinder head 33.
20

exhaust valve 34,
port 35,

An inwardly opening poppet

disposed in the inlet 31 of the exhaust

controls fluid communication between the expansion

cylinder 14 and the exhaust port 35.
Typically,
engine
25

such

as

in

that

a

naturally

shown

in

aspirated

FIG.

1,

the

split-cycle
compression

cylinder displacement volume (Vd) required to intake a given
charge

(or

mass)

of

air

is

largerthantherequired

displacement volume of a cylinder of a conventional engine
for the same charge of air.

The compression cylinder of anaturally aspirated split-cycle engine must be made larger
30

because during engine operation thereisalwaysatrappedcompressed air mass present in the compression cylinder atthe end of the compression stroke.
subsequentintakestrokeTherefore,

immediately

during thefollowingthewo 2012/040431

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7
compression

stroke,intakeair

cannot

be

drawn

into

the

compression cylinder until the compression piston drops down
far enough from top dead center so that the pressure of thetrapped air mass
5

is

equal

to

atmospheric pressure.

Thus,part of the volume swept by the compression cylinder duringtheintake stroke

is not utilized for

air intake.

Hence,

the compression cylinder must be made larger so that it has
sufficient volume to draw a necessary amount of intake air
during the remainder of the intake stroke.
10

displaced volume
naturally(ordecreases

aspirated

the power density of

split-cycle

specific power)

This increase inbeing

engine,defined

as

the

the

a

typicalpower

densitybrake

power perengine displacement, usually expressed as kilowatts/liter orhorsepower/liter.
15

It is also known in the art of internal combustion
enginestooperateaconventional

engine using

the Miller

cycle.

The efficiency of an internal combustion engine isincreased if the gas is expanded more during the expansion
stroke than it is compressed during the compression stroke.
20

In

the

Miller

cycle

of

a

conventional

engine,thisistypically accomplished by early or late inlet valve closing(IVC),

which

decreasestheeffective

relative to the expansion ratio.

compression

For example,

valve of a conventional engine is closed late25

thecompression

strokethat

follows

the

ratio

if the inlet(i.e.,

intakeduringstroke),aportion of the intake air that was drawn into the cylinder
during the intake stroke is pushed back out of the cylinder
through the intake port.

The intake valve may be kept open

during about the first 20 percent of the compression stroke.
30

Therefore,

actual compression only occurs in about the last

80 percent of the compression stroke.
Referring to FIG. 2A, an exemplary embodiment of apressure vs.

volume(PV)diagram ofanaturally aspirated

wo 2012/040431

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8

engine utilizing late IVC to effect Miller cycle operation
isshown.

Thoughaspirated engine,

thisembodiment

depictsanaturally

it is know that the same principles apply

to turbocharged engines as well.
5

As shown in FIG.

2A,

during the intake stroke ofthe piston from TDC to BDC, the cylinder pressure follows aconstant

pressure

line

from

point

finallyto5.

Duringthestroke,while

subsequent
10

point

compression

6

throughinitial
the

point

1

and

portion

of

the

valve

is

intakeleft open the cylinder pressure retraces the pressure linefrom point 5 back to point 1.

Then,

at point 1 the intake

valve closes and the cylinder pressure increases from point
1 to point 2 during the remainder of the compression stroke.
The
15volumeswept

bycanceled by the volume

thepiston

swept

along

the

path

along the path 5-1,

1-5

is

and theeffective compression ratio is the volume at point 1 divided
by the volume at point 2 rather than the volume at point 5
divided by the volume at point 2 for the Otto cycle.
Referring
20

to

FIG.

2B,

the

effectcanbeachieved in the Miller cycle by early inlet valve closing.
In

this

case,

the

pressureremains

intake stroke from point 6 to point 1.
intakevalve

closes,andthecompression stroke,

constant

duringtheThen at point 1 thepressure

decreases from point 1 to point 7.
25

same

in

the

cylinderDuring the subsequent

the pressure increases from point 7 topoint 1, canceling the previously traced path, and continues
to point 2 during the remainder of the compression stroke.
The

net

That is,30

usedresultisthesame aslate

intake valve closing.

less than the entire piston stroke is effectively

forcompression

compression,
ratio

for

therebydecreasing

increasedtheefficiency

decreasing the mass of charge air per cycle.

effective
while

also

wo 2012/040431

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9

The

increaseinefficiency

of

the

Millercycle(typically 10 to 15 percent greater than the Otto cycle)

isnegatively offset by a decrease in indicated mean effective
pressure (IMEP) and power density that is a result of a loss
5ofchargevolumeinairbecause

only

part

of

the

cylinderisfilledthewithdisplacement volume is sacrificed).

totalchargedisplacedairAccordingly, to achieve

the same amount of power as an Otto cycle engine,
cycle engine typically must be made larger,
10(i.e.,

a Milleror boosted,orboosted more aggresively.
Additionally, the later or earlier IVC occurs, thefaster
the

air

This
15

the piston isistherefore,

thefasterflowing over theintake valve whentosignificant

pumpingloses,

which

engineefficiency

(i.e,

brake

specific

leads

reducestraveling and,it closes.
greatly
fuelconsumption (BSFC)) of conventional Miller cycle engines.
SUMMARY OF THE INVENTIONThe

Scuderiengineisasplit-cycle

designthat

divides the four strokes of a conventional Otto cycle over
20

(in

its

pair.

simplestThecylindercompressor
andreciprocates25

configuration)
includesassociatedacompressor

and expander

at

least

compression

onecompression

which

compression

strokes.

throughtheintake

The expander includes

atleast one expansion cylinder andassociated expansion piston,
expansion(or power)

andpiston,which reciprocates through theand exhaust strokes.

The expander andcompressor are connected by a crossover passage.
Millerinternal
30

cycle

combustion

operationenginesplit-cycle or conventional,
configured such that
expansion

strokecompression stroke.

the

than

of

(RICE),

anywhetherreciprocating
the

RICE

is

necessitates the engine to be

gas

is

expanded moreduring

the

it

is

compressed

duringtheAccordingly,

gas will be over-expandedwo 2012/040431

PCT/US2011/052711

10during5

theexpansion

stroke,efficiency(i.e.,

consumption(BSFC))

operation

typically accomplishedintake

is

valve

a

resultingdecreaseIn

closing

in

a

in

an

brake

fuel

RICE,

Millerthrough early or

However,in

specific

conventional(IVC).

increase

in

a

late

split-cycle

engine, Miller cycle operation may be achieved by physicallyreducing

thecompressor

displacement

relativetotheexpander displacement.
A Miller
10

cycle

configuration

of

the

split-cycle

engine was constructed with a turbocharged intake manifold
connected

toadownsized

split-cycle

engine's

relative to the split-cycle engine's expander,

compressor

and has beenmodeled in 1-dimensional cycle simulation software.
positive
15

interactions

were

found

betweentheengine and Miller cycle operating principles.
1.

Downsizing

thecompressor

Severalsplit-cycle

Namely:relativetotheexpander to achieve Miller operation in a split-cycle engine
does

not

achieve
20

waste
the

part

of

the

compressor's

reduced compression.

displacement

Contrarily,toin order toconfigure a conventional RICE for Miller operation at fullload,

its cylinders must be sized for the full displacementof its expansion stroke and part of that displacement must

25

necessarilybeduringcompression

thewastedtoaccommodate
stroke.

early

The

or

late

IVC

of

the

reduction

compressor displacement on the split-cycle engine provides a
generally

higher

brake

mean

effective

pressure(BMEP)

compared to a conventional RICE with Miller cycle operation.
2. Reduction of the compressor displacement allows
Miller cycle operation while still closing the intake valve30

at or near an optimum trapped mass condition
volumetricefficiency) .

This

results

in(i.e.,

optimum

reducedpumping

work compared to the Miller cycle applied to a conventionalwo 2012/040431

PCT/US2011/052711

11

RICE,

due

to

the

avoidanceofduring a period of high piston
3.

5

Whetherratio

expansion during theof gasofduringthecompression

of

work

load

compressor.

Since

the

turbo

via

the

turbine,inincrease

terms
in

aftercooling,valve

split-cycleorfactor(i.e.,

expansion

stroketogasstroke)determines

the

that

compressor

gases

intakeMillercompression

engine's

exhaustisthesplit-cycle"cost"

isshiftedtothecompressor
it

is

from

turbocharger
drivencompresses

by

at

a

of

crankshaft

energy,providing for

work

deliveredtocrankshaft.

the

the

the

lower
a

net
With

this also allows cooling later in the overall

compression process,resulting in a lower end-of-compression

temperaturesame

aspect
for

of

for
the

increasing

and

hence

the
lower

pressure.

The

end-of-compression

compression

Miller factors,

20

RICE

the

level

15

the

magnitude

the(and therefore air) velocity.

conventional,

compression

10

closingpressuresknock

suppressing

temperature
at

more

allowsaggressive

resulting in higher total air and fuel flow,

power.

The

previously

mentioned

split-cycle's

ability to utilize moreaggressive Miller factor

operation

thanconventional

allowsenginetakegreaterRICEs

advantage

of

the

these

split-cycle

characteristics

of

to
the

Miller cycle.
4.
25

For a

given boost level,

constraints used by the engine(i.e.,

expanderthecylinderpressureorwithin the operatingknock fraction,like),

thepeak

split-cycle

engine is most efficient when volumetric efficiency of boththecompressor

Typically,
30

the

and

expandervolumetricareefficiency

measured relative

tobemore preferably at

at

least

0. 75,

the

of

or

near

the

compressor

intake manifold conditions

preferably at least 0.85,
Typically,atvolumetric

least

optimum.
isand should

0. 8,even more

and most preferably at least 0.9.
efficiency

of

the

expanderiswo 2012/040431

PCT/US2011/052711

12

measured
least

relative

0.90,

to

more

ambient

conditions

preferably

at

and

least

should be

1.0,

even

at

more

preferably at least 1.1, and most preferably at least 1.2.
More particularly,
5

theinventionain an exemplary embodiment ofsplit-cycle

engineincludesacrankshaft

rotatable about a crankshaft axis and an expander having an
expanderdisplacement

expansion
cylinder10pistonvolume.

slidablyThe

expanderreceived

includeswithinand operatively connected to

the

an

an

expansion

crankshaft

suchthat the expansion piston reciprocates through an expansion
stroke and an exhaust stroke during a single rotation of the
crankshaft.

The

split-cycle

enginealso

includesacompressor having a compressor displacement volume that is
90 percent or less
15

The

compressor

received

than the expander displacement volume.

includeswithinaacompression

compression

pistoncylinderandslidably

operatively

connected to the crankshaft such that the compression piston
reciprocatesstroke20

throughduringaanintakesinglestrokerotation

of

and
the

a

compression

crankshaft.

Acrossover passage interconnects the compression cylinder andthe expansion cylinder.
least
intakeacrossover

The crossover passage includes at

expansion

manifold connects

to

valve
the

disposed

therein.

An

compression

cylinder.

Aboosting device is connected to the intake manifold and is25

operabletoprovidea

1. 7 barabsolutepressure level to the intake manifold.
disposed

betweenthecylinderandoperable

therebetween.
30

event

which

is

intaketomanifold
controlorgreaterboost

An intake valve is
and

the

fluid

compression

communication

The intake valve has an intake valve closing
is

timed

to

provideacompression

cylindervolumetric efficiency relative to intake manifold conditionsof 0. 7 5 or greater.

The compressor displacement volume issized relative to the expander displacement volume such that

wo 2012/040431

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13

the combination of compressor displacement volume and boostpressurelevel

provideanexpandervolumetricefficiency

relative to ambient conditions that is 0.90 or greater.
The boost operating device may be a turbocharger.
5The

compressor

mayincludeapluralityoftheexpandermay

includeacylinders,andexpansion

cylinders.

includeacrossovercrossover passage,
10

and theThe

split-cycle

compression

compression
plurality

engine

valve

of

may

also

in

the

disposed

wherein the crossover compression valve

crossover expansion valve

form a pressure chamber

therebetween.
The

compressor

percent or less,less15

displacement

75 percent or less,volume

may

be

80

or even 70 percent orthan the expander displacement volume.

The boosting

device may be operable to provide boost pressure levels of
2.0 bar absolute or greater, 2.3 bar absolute or greater, oreven 2.5 bar absolute or greater.
efficiency

relative

0. 80 or greater,
20

tointake

The compressor volumetricmanifold

0. 85 or greater,

conditions

may

be

or even 0. 90 or greater,while the expander volumetric efficiency relative to ambientconditions may be 1. 0 or greater,

1. 1 or greater,

or even

1.2 or greater.
A method of operating a split-cycle engine is also
disclosed.
25

The

split-cycle

engineincludesacrankshaft

rotatable about a crankshaft axis and an expander having an
expanderdisplacement

expansion
cylinderpistonvolume.

slidablyThe

received

expanderwithinand operatively connected totheincludesananexpansion

crankshaft

suchthat the expansion piston reciprocates through an expansion
30

stroke and an exhaust stroke during a single rotation of the
crankshaft.

The

compressor

havingasplit-cycle

enginealso

compressor

displacement

includesvolume.

aThe

compressor includes a compression piston slidably received

wo 2012/040431

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14
within acompression cylinder and operatively connected tothe crankshaft such that the compression piston reciprocatesthrough an intake stroke and a compression stroke during asingle5rotationinterconnects

of

the

the

crankshaft.

compression

cylinder.

The

crossovercrossover

expansionvalve

A crossover

cylinderandthepassageincludesdisposed

therein.

passage

expansion

at

least
An

manifold connects to the compression cylinder.

aintakeA boosting

device is connected to the intake manifold and is operable
10

to provide a boost pressure

level

to

the

intake manifold.

An intake valve is disposed between the intake manifold andthecompression

cylinder and iscommunication therebetween.
the present
15

operabletocontrolfluid

The method in accordance withinvention includes

the following

steps:

sizingthe compressor displacement volume to be 90 percent or lessthan the expander displacement volume; delivering a 1.7 bar
absoluteorgreaterboostpressure

level

to

the

intakemanifold from the boosting device; timing the closing of theintake
20

valve

such

that

a

compressor

volumetric

relative to intake manifold conditions isandsizingthecompressor

displacement

the expander displacement volume

0. 75 or greater;

volumesuch that

efficiency

relativethe combinationof compressor displacement volume and boost pressure
provide25

anexpandervolumetrictoefficiency

level

relativetoambient conditions of 0.90 or greater.
These
invention will

and

other

be morefeaturesandadvantages

fully understood from the

of

the

followingdetailed description of the invention taken together with theaccompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS30

In the drawings:
FIG.

1

is

split-cycle engine;across-sectional

view

of

a

prior

art

wo 2012/040431

PCT/US2011/052711

15

FIG.

2 is a graphical illustration of a prior artpressure vs. volume diagram for a conventional engine Miller
cycle implemented with early and late intake valve closing;FIG.
5

3 is a transverse cross-sectional view of an

exemplary embodiment of a turbocharged split-cycle engine inaccordance with the present invention;
FIG. 4 is a cross-sectional top view of the splitcycle

enginetaken

through

line

3-3

of

FIG

2

with

fuel

illustration

of

brake

brake

effective

injectors superimposed;
10

FIG.

5

is

a

graphical

specific

fuelconsumption

pressure(BMEP)

and(BSFC),

volumetricefficiency

mean(Vol Eff)asafunction of compressor stroke length at a fixed boost of 1.7
bar absolute;
15

FIG.

6 is

a

graphical

efficiency ofthecompressor

cycle

as

functionengineaillustration of volumetricand theofexpandercompressor

boost levels of 1.7 bar absolute,ofstrokeasplit-

length at

2.0 bar absolute,and 2.3

bar absolute, at full engine load and an engine speed of 4000
20rpm;
FIG. 7 is a graphical illustration of a comparisonoftemperaturetemperatureandpressure

and pressure

for

for

pistonturbo

compression

versuscompression

followed byillustrationofpiston compression;
25

FIG.

8

is

a

graphical

turbocharger efficiency assumptions

as

level

turbochargers

for

high

and

low

efficiencyafunctionoverallof boost

atenginespeeds of 1400 rpm (low speed) and 4000 rpm (high speed);
FIG.
30

9 is

effective pressure(BSFC)asaa graphicalillustration of brake mean

(BMEP) and brake specific fuel consumptionfunctioncompressor stroke length)

of

boost

level

(and

corresponding

for the low efficiency turbocharger

wo 2012/040431

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16
and the high efficiency turbocharger of FIG.

8 at an engine

speed of 1400 rpm;
FIG.

10 is a graphical illustration of brake meaneffective pressure
5(BSFC)

asa(BMEP) and brake specific fuel consumptionfunctionofcompressor stroke length)

boost

level

(and

corresponding

for the low efficiency turbocharger

and the high efficiency turbocharger of FIG.

8 at an engine

speed of 4000 rpm; andFIG.
10effective

11 is a graphical illustration of brake meanpressurepressure,(BMEP),

compression

cylinderpeak

and expansion cylinder peak pressure as a functionof engine speed at a boost of less than 3 bar absolute and atfull engine load.

DETAILED DESCRIPTION OF THE INVENTION
15

The following glossary of definitions of acronyms
and terms used herein is provided for reference and clarity.
Unit of pressure, 1 bar = 10 5 N/m 2 •

Bar:

20

25

Bar-abs(Bar-absolute) :Pressure referenced to a

vacuum,i.e.,

atmosphericstandard

equals

1.013

bar-absolute.
BMEP:

Brake

mean

effectivereferstotheoutputas

outputshaft),

after

frictionpressure.

delivered

The

to

the

term

and parasitic

lossesBrake Mean Effective Pressurethetorqueengine'sbrake

output(MEP)

expressed

value.

"Brake"crankshaft

are accounted for.

mean effective pressure

in

(or

(FMEP)(BMEP)

terms

of

is
a

BMEP is proportional

to the brake torque divided by engine displacement.

This isthedueperformance

friction.
30

pressurecompletecaseisparameter

Accordingly,
usually

also

taken

after

BMEP=IMEP- FMEP.

expressedinatypicalfrictionvalue

lossesFriction,terms

known as Friction Mean Effective Pressure
definition,

the

of

an

to

in this

MEP

value

(or FMEP). In this(loss)

would

be

a

wo 2012/040431

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17

positive

value

of

FMEP;

in

other

words,

apositive

FMEP

would be a net reduction of work provided to the crankshaft.
Boost
from
S

or Boostaboosting

Pressure

Level:

device,such

Intakeasamanifold pressure

turbocharger,

in

bar-

absolute.
Brake Specific Fuel Consumption(BSFC):

A measure of fuelefficiency within a reciprocating engine.

It is the rate offuel consumption divided by the brake power produced and istypically
10

expressed(g/ (kW·h))

in

units

of

grams

per

kilowatt-hour

BSFC may also be thought of as power-specificfuel consumption.
Compressorofcompression
compressed
15

duringaasplit-cycle engine:

cylindersairnumberofpistons

providing

expansion

cylinder

least

one

commoncompression

strokeforanengine.

total

associatedatsplit-cycle

to

and

The

Typically,aoperatingcompressor

cycle
of

a

of

a

split-

cycle engine includes a single compression cylinder and its
associated compression piston.

However,

it is possible for

a compressor to include a plurality of compression cylinders
20

and pistons.
Compressor displacement volume of a split-cycle engine:
total combined displacement volume

(Vd)

The

of each compression

cylinder of a compressor.

25

Displacement

volumeofacylinder(Vd),

conventionalorsplit-cycle:

Thevolumedisplacedbypistonaawithincylinder of an engine as the piston travels from its bottom
dead center position to its top dead center position.
Effective compression ratio of a conventional engine:

30

ratio

of

the

enginewhen

volumethewithinacylindercylinder's piston

is

at

of

The

a

conventionalits

intake valve

closing position to the volume within the cylinder when thecylinder's piston is at its top dead center position.

wo 2012/040431

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18

Expanderof

expansionasplit-cycle

cylindersengine:andThe

associatedtotal

numberpistonsofreceiving

compressed air from at least one common compression cylinder
during5

anexpansion

split-cycle engine.
engineincludesstrokeforanoperating

cycle

of

a

Typically, an expander of a split-cycleasingleassociated expansion piston.

expansion

cylinderandits

However, it is possible for an

expander to include a plurality of expansion cylinders andpistons.
10

Expander displacement volume of a split-cycle engine:
total

combined displacement

volumeThe(Vd)of each expansion

Friction mean effective pressure.

Theoretical meancylinder of an expander.
FMEP:

effective pressure required to overcome engine friction plus
15

parasiticpressure

losses.
lost

due

Can
to

be

thought

friction.

of

Sign

as

mean

convention

effective
for

this

discussion is that a friction loss is a positive FMEP value.
Full(100%)Engine Load:

The maximum torque that an engine

can produce at a given speed.
20

Geometric compression ratio of a conventional engine:The

ratio of volume within a cylinder of a conventional engine
when

the

cylinder'spositiontothepistonvolumeisatwithinitsbottom deadthecylindercenterwhen

thecylinder's piston is at its top dead center position.
25

Geometric expansion ratio of a conventional engine:
conventional engine,strokes

are

For a

wherein the compression and expansion

contained

withinthesamecylinder,thecompression ratio and the expansion ratio are identical.
Indicated:
30

Refers to the output as delivered to the top of

the piston, before friction losses are accounted for.
Indicated Mean Effective

Pressure(IMEP):

Mean effective

pressure calculated from in-cylinder pressure.

IMEP is thewo 2012/040431

PCT/US2011/052711

19

integration of the area inside the P-V curve, which is also
proportional

to

the

displacement

volume.

indicatedIn

enginefact,

alltorquedividedbyindicatedtorqueandpower values are derivatives of this parameter.
5

IVC:

Intake valve closing or intake valve closing event.

The timing of the closing of the intake valve relative tobottom dead centerintake

valve

(BDC)

closing

isof the compression piston.

Early

the

valve

closingoftheintakebefore BDC and late intake valve closing is the closing of
10

the intake valve after BDC.
Knock:The tendency of an air/fuel mixture to self-ignite

during compression.
Knock Fraction(also commonly referred to as Knock Index orKnock Margin or Knock Number) :15

providesarelative

particular

air/fuelcompression.
value

20

A predicted parameter whichindication

of

the

tendencyofamixturetoreachself-ignition

duringSelf-ignitionisusually denoted by a

knock

no

tendency

fraction

of

1

while

to

self-ignite

isusually denoted by a knock fraction of zero.

For example, aknockphysicalandself-ignition

have

fractionchemicalof

0.8

conditions

indicates

prerequisite

that
for

the

reached 80% of the value required to generate self-ignition.
mm:

25

Millimeter.

Parasitic Losses:

Energy losses

incurred by anoperateaccessories

andancillarydevices.

energy

to

items

such

pumps

drive

generating devices.

asengine

This
and

to

includeselectricity

The brake output of an engine is reducedby these losses.
Power
30

Densitygenerated

at(orSpecific

Power):

thecrankshaft)

per

Brake
enginepower(powerdisplacement,

usually expressed as kilowatts/liter or horsepower/liter.

wo 2012/040431

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20
RICE:

Reciprocating internal combustion engine.

Any engine

havingreciprocating

combustion

process.

pistonsandaninternal

May be either a conventional engine or a

split-

cycle engine.
5

RPM:

Revolutions per minute.

Split-Cycle

EngineDisplacement

Volume:compressor and expander displacement

The

volumes

sum

of

the

split-

ofthecycle engine.
Turbo:
10

Turbocharger.

Volumetric Efficiency of a cylinder,
cycle:

Abbreviated Vol Eff.

conventional or split-

The ratio of the actual mass of

air trapped in the cylinder during the closed portion of thecycle, compared to the theoretical mass of air that would betrapped in the cylinder's displaced volume if it were filled
15

with

air

of

referenceeithersame

location.

ambient

pressure)

20

the

or

(i.e.,
intakepressureandtemperature

locationThe

referenceair

atatmospherictemperature

manifold

conditions(air

manifoldtemperatureand pressure).

may

definedthefollowing

equation:

is

the

of

be

ma/(da*Vd),

by

where

rna

mass

is

as

Volumetric

air

the

typicallyatandintakeEfficiency
Vol Eff

trappedinthecylinder per cycle, da is the density of air at the reference
state, and Vd is the cylinder displacement volume as defined
herein.
25

Volumetricengine:Efficiencyofacompressorofasplit-cycle

The ratio of the actual mass of air trapped in thecompressor during the closed portion of the cycle,

comparedto the theoretical mass of air that would be trapped in thecompressor's displaced volume if it were filled with air of30

the same pressure and temperature as the reference location.
Volumetric Efficiency of an expander of a split-cycle engine:
The ratio of the actual mass of air trapped in the expander
duringtheclosedportionofthecycle,

comparedtothewo 2012/040431

PCT/US2011/052711

21
theoretical

mass

of

air

that

would

be

trappedintheexpander's displaced volume if it were filled with air of thesame pressure and temperature as the reference location.
Xover (Xovr):
5

Crossover.

Referring now to FIGS.
detail,numeralembodiment
presentof50

a

3 and 4 of the drawings in

generally

split-cycle

invention.

indicates

engineEngine50

inanexemplaryaccordance

includes

a

with

the

crankshaft

52

rotatable about a crankshaft axis 54 in a clockwise direction
10

as shown in the drawing.
angularly displaced

The crankshaft 52 includes adjacentleadingandfollowing

58, connected to connecting rods 60,

defining
15

Engine50

furtheraofadjacent

pair

crank

throws56,

62, respectively.

includesacylindercylinders,

in

block

64

particularacompression cylinder 66 and an expansion cylinder 68 closed
by a cylinder head 70 at one end of the cylinders opposite
the crankshaft 52.
A compression piston 72 is received in compression
cylinder 66

20(thereby defining a compressor

(66,

72))

and isconnected to the connecting rod 62 for reciprocation of thepiston between top dead center(BDC)

positions.

An

expansion cylinder 68(TDC)

expansionand bottom dead center

piston74

isreceived(thereby defining an expanderin(68, 74))

and is connected to the connecting rod 60 for similar TDC/BDC
25

reciprocation.
leads
In

In this embodiment,

the expansion piston 74

the compression piston 72 by 20

otherpositionwords,
20

thedegreesdegrees

compression pistonofcrankshaft

72

crank angle.

reachesrotation

its

TDC

after

the

expansion piston 74 reaches its TDC position.
30

While
illustrates atheembodiment

shown

in

the

drawings

compressor that includes only one compression

cylinder connected by a crossover passage to an expander that
includes only one expansion cylinder,

it should be understoodwo 2012/040431

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22that thecompressor may include morecylinderassociated

and/orthecylinder5expander

with

and

may

includethan oneconnected
more

to

the

expander,

than

one

expansionassociated with and connected to

However,fordescription

purposesrefers

of

to

the

compressor.

thefollowing

simplicity,acompressor

compression

havingasinglecompression cylinder paired with an expander having a single
expansion cylinder.
As explained in more detail below, the displacement
10

volumethe(Vd)of the compressor is significantly smaller thandisplacement

specifically,
percent,

(Vd)

of

the

the displacement of the

80 percent,displacement
15

volumeof

75 percent,theexpander.

Morecompressor may be

90

70 percent or less than the

expander.

Accordingly,

the

gas

is

expanded more on the expansion stroke than it is compressedon the compression stroke to enable Miller cycle operation.
Thecylinder head70

providesthestructuregas flow into, out of and between the cylinders 66,
the order of gas flow,20

manifold

(or

68.

for
In

the cylinder head includes an intake

intake port)

76

through which

drawn into the compression cylinder 66,
crossover (Xovr) passages

intakeair

isa pair of separate(or ports) 78 and 79 through which

compressed air is transferred from the compression cylinder
66
25

to

the

expansion

cylinder

68,

and

an

exhaustport

80

through which spent gases are discharged from the expansion
cylinder.
Gas

flow

into

the

compression

cylinder66iscontrolled by an inwardly opening poppet type intake valve
82.
30

79

Gas flow into and out of each crossover passage 78 andisvalves,inlet

controlled
i.e.,
ends

expansion

by

a

pair

crossoverof(XovrE)

the

Xovr

of

outwardly

compression
passages(XovrC)78,

79

opening
valvesandpoppet86

atcrossovervalves 88 at outlet ends of the crossoverwo 2012/040431

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23other ignition devices
intheexpansion

chargemay

belocated at appropriate locations

cylinderignitedwherein

a

mixed

and

burnedalso

includesfuel

during

the

and

air

expansion

stroke.
15

The
such

as

engineturbocharger

50

100,

a

a

boostingsupercharger,ordevice,similar,capable of raising cylinder intake charge pressures up toand beyond 1.7 bar absolute (such as 2 bar absolute, 2.3 barabsolute, 2.5 bar absolute 3 bar absolute or more).
20

In thisembodiment, turbocharger 100 includes an exhaust turbine 102
driving a rotary compressor 104.

The turbine has an exhaustgas inlet 106 connected to receive pressurized exhaust gas
from the exhaust port 80 of the engine 50.

The turbine 102

utilizes the waste energy of the exhaust gas to drive the25compressor 104,

which draws

in ambient air at

atmosphericpressure through an air inlet 108 and discharges pressurized
air through a compressed air outlet 110.

The compressed air

passes through an intercooler 112 and enters the air intake
manifold 76 of the compression cylinder
30

66 at an absolutepressure boosted above that of atmospheric pressure.
Although the turbocharger 100 is shown as a single
stage

devicecoupledwithanintercooler,itshouldbeunderstood that other turbine systems are within the scopewo 2012/040431

PCT/US2011/052711

24

of

the

invention.

multiplestagesThus,ofthe boosting device may

turbocharging,

and

the

devicecoupled with multiple stages of intercooling.
such turbocharger systems include,5

includemaybe

Examples ofbut are not limited to,two stage systems, parallel sequential systems or modulating
seriessequential

systems.

Theboosting

deviceisnotparticularly limited to a specific structure as long as theboosting device is capable of increasing the pressure of theintake air to a certain amount above atmospheric pressure.
10

In

the

present

split-cycle

engine50,

theturbocharger 100 uses waste heat energy(energy that would

otherwise be dispensed to the atmosphere)

to do the initial

work of compression.

This reduces the total work requiredby the compression piston 72 of the engine 50 to reach the
15

same

end

pressure,ofcompression

temperature,

constraints(e.g.,

knock fraction,peakcylinderor the like)that anaturally aspirated split-cycle engine would achieve for agiven charge of air.
Accordingly,
20

while operating the engine 50 within

its predetermined constraint(peak cylinder pressure,knock

fraction or the like) limits, the greater the boost from theturbocharger system 100 the more the total compression work
load

can

be

shiftedfromtheturbocharger compressor 104.
25

compressor

66

relativecompressor

As a result,

to

the

66,

72

to

the

the size of the

expander

68

can

be

progressively reduced with increasing boost pressure without
exceedingtheengine's

downsizing

of

the

operatingcompressor

relative

greatly increases the Miller factor
30

expansion

duringtheexpansion

constraints.
to(i.e.,

stroke

to

the

This
expanderthe ratio of gas
gas

compression

during the compression stroke) and therefore increases brakemean effective pressure
fuel consumption (BSFC) .

(BMEP)and decreases brake specificwo 2012/040431

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25

Reduction of the compressor displacement relative
to the expander displacement allows Miller cycle operation
while

still

optimum
5

closingtrappedthemass

intake

valve

82

at

or

condition(i.e.,

optimumnear

ancompressor

volumetric efficiency relative to intake manifold conditionsof 75%,

80%,

85%,

cylinder66

to90% or more)
effectMiller

Downsizing the compression
cycle

operation

resultsin

reduced pumping work compared to a conventional Miller cycle
engine, which utilizes early or late intake valve closing to10

implement

the

same

Millercycle

operation.

Reductioninpumping work is due to the avoidance of closing the intake

15

valve 82 during a period of high piston 72(and therefore

air)velocity50

cycle.

By

in

the

contrast,

split-cycle
pistonengine'svelocity

andoperatingpumping

losseswould increase significantly in a conventional engine as theintake valve closes progressively later or earlier (i.e., ascompressor volumetric efficiency gets
for Miller operation.

As a result,progressively worse)the split-cycle engine

50 can achieve more aggressive Miller factor operation and20

benefits than a conventional engine.
Additionally, downsizing the compressor to achieve
Miller operation in a split-cycle engine does not waste partof the compressor's displacement.

Contrarily,

in order toconfigure a conventional engine for Miller operation at full
25

load,

its cylinders must be sized for the full displacementof its expansion stroke and part of that displacement mustnecessarily be wasted toaccommodate early or latevalve closing during the compression stroke.

intakeAs a result,

the reduction of the compressor displacement on the split30

cycle

engineprovideseffective pressure(BMEP)agenerally

higher

brake

mean

compared to a conventional engine

with Miller cycle operation.

wo 2012/040431

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26
As

illustrated in FIG.

5,

care must be taken to

match the boost to the compressor displacement for optimum
performance.

Insufficient

displacement
5

willresultinboost

and/orinsufficientcompressor

fillingof

expander, as shown towards the left side of FIG. 5.
been

found

becomes

that

when

undesirably

the

low

expander(i.e.,

volumetric

below

0.90

the

It hasefficiency

whenmeasured

relative to ambient conditions), performance is non-optimum.
If
10

the

expanderindicatesthat

displacement
performance
FIG.

15

5,

volumetricthevolume
will

for

a

compressor
tosuffer.
given

efficiency

is

too

not

have

enoughexpanderandengine

did

fill

the

More

particularly,

boost

level,

withinconstraints used by the engine,

the

most

efficiency

efficientwhenvolumetric

low,

as

shown

the

it

in

operatingsplit-cycle engine
of

both

is
the

compressor and expander are at or near optimum.
In FIG. 5, the turbocharger 100 was set to provide
a constant boost of 1.7 bar absolute to the intake manifold
76 of engine 50, and the intake valve 82 closing (IVC) timing
20

wasutilizedpressure

to

(PCP)

prevent

theconstraint72)

and expander

diameter of 87 mm,
25

peak

cylinderof the expansion cylinder 68 from exceeding apredetermined maximum value.
(66,

of

Additionally,

thecompressor(68, 74) were set to have a fixed bore

and the expander was also set to have afixed stroke length of 87 mm.

The compressor stroke length

was varied from 50 mm to 90 mm (i.e., the displacement volumeofthecompressor

wasvaried by changing

the

size

of

the

stroke length of the compression piston 72).
The
30

top

graph

specific fuel consumption
stroke length.
brake

mean

120

of

(BSFC)

FIG.

5illustrates

as a function of compressor

The middle graph 122 of FIG.

effective

pressurecompressor stroke length.

brake(BMEP)

as

5 illustratesafunctionofThe bottom two graphs 124, 126 ofwo 2012/040431

PCT/US2011/052711

27
FIG.

5

5

illustrate

volumetriclength,efficiency

wherein

the

as

a

functioncompressor

strokeefficiency

126

is

referenced

relativetoconditions

and

the

expandervolumetricefficiency

ofcompressor volumetricintakemanifold124

isreferenced relative to ambient conditions.
At a 50 mm compressor stroke length, the compressor
(66,

72)

displacement

volume

mm/87

mm

0.57)

the

volume.
10

the

At

intakevolumetric

of

only

expanderthisdownsized

valve

82

can

efficiency

is57

percent(68,

74)

compressor
be

while

15

cylinder 68.

That is,an optimum(or maximum)

displacement
volume,tomaximum

provideengineoperate well below the peak pressure

50

displacement

operated
the(i.e.,

50

limit

continues

of

the

to

expanderintake valve 82 is timed to close attrapped mass

compressor volumetric efficiency is

condition,

wherein theapproximately

referenced to intake manifold 7 6 conditions

0. 92

when

(the theoreticalvolumetric efficiency limit being 1.00).
However,a20

maximum volumetric

sized

too

small

boost level.
expander25

at a 50 mm stroke,tosufficiently

As a

is

efficiency of

result,even when operating at

0. 92,

fill

the

the

compressor

expanderthisthe volumetric efficiency of theunacceptably

low,

value

of

referenced

toconditions(whereinanvolumetricefficiency

of

less

0.90

undesirable).

Accordingly

andabout 10.8 bar(graph 122)

ambientatishaving

a

than

is

undesirably,

BMEP

0. 85

when

expanderconsidered
drops

to

and BSFC rises to about 285 g/kw-

hr (graph 120).

increases
30

Referring

to

FIG.

from

to

65

50

5,

mm,

as

compressor

intakevalve

82

strokelengthcontinues

toclose at optimum trapped mass conditions and the compressor
continues

tooperate0. 92.

Additionally,

50

65

to

mm,

the

at

a

maximum volumetric efficiency ofas the compressor stroke increases fromexpandervolumetricefficiency

increases

wo 2012/040431

PCT/US2011/052711

28
from 0.85 to 1.10 as the expander receives greater amounts ofcharged air from the compressor.
expansion

cylinder68

The peak pressure of thecontinuously

rises

as

the

expandervolumetric efficiency increases until the predetermined limit
5

of

peak

pressureisreachedwhen

theexpandervolumetricefficiency reaches 1.10.
Atacompressor

strokelengthofcompressor displacement volume is 75 percent
mm
10

=

0. 7 5)

ratio

of

of

the

expander displacement

compressor

toexpander65mm,(i.e.,

65 mm/87

volume.

volume,

the

the

At

this

engine50

operating conditions are as follows:
1) Boost Pressure is 1.7 bar absolute;2) Compressor volumetric efficiency is maximized at 0.92
relative to intake manifold conditions;15

3)

Expander volumetric efficiency is maximized at 1.10

relative to ambient conditions; and4)

Expanderpeakpressure

is

at

its

predeterminedmaximum limit.
Essentially,
20

while

for the given boost level of 1.7 bar absolute,

operatingwithinthecylinder pressure limit,
to provide
0.92

25

an

relative

predetermined

expansion

the intake valve 82 has been timed

optimum compressor volumetric efficiency oftointakemanifold

conditions,compressor

displacement

volumepercent

thedisplacement

of

peak

expander

has

been

sized

volume

to

and
to

the

be

provide75amaximum expander volumetric efficiency of 1.10 relative toambient

conditions.

Under

these

optimized

parameters,

Miller cycle operation is most beneficial and the engine 50
operates at a peak BMEP of 12.8 bar and a minimum BSFC of30

282 g/kw-hr.
It should be noted that the split-cycle engine 50(and split-cycle engines in general)can function with lessthan ideal operating parameters and still derive significant

wo 2012/040431

PCT/US2011/052711

29
benefits

fromMillercycle

operation.

benefitsofMillercycle

operation

However,
to

have

for

the

practical

significance on a split-cycle engine the following thresholdengine operating parameters should be met or exceeded:
51) The boosting device is operable to provide a 1.7 bar
absolute or greater boost pressure level to the intake
manifold;2) The intake valve closing event is timed to provide acompressor

10volumetricefficiency

relativetointakemanifold conditions of 0.75 or greater; and
3) The compressor displacement volume is sized relative
totheexpanderdisplacement

volume

such

that

the

combination of compressor displacement volume and boostpressure
15

levelprovidesan

expandervolumetricefficiency relative to ambient conditions that is 0.90
or greater.
It

20shouldalso

volumetricefficiency

isparameters,

the boost pressure level to the intake manifold

and

the

expandercompressor

of

noted

boost

that

dependent

displacement

displacement

combinationbeprimarily

volumevolume.

Itpressure

theexpanderupon

relative
is

level

two

to

the

primarily

the

and

compressor

displacement that provides the proper air mass to fill the
expander25

andattaintherequired

expandervolumetricefficiency.
Referring again to FIG.

5,

for compressor strokes

greater than 65 mm, the compressor becomes too large relative
to the expander to maintain the maximum compressor volumetricefficiency of 0. 92 without exceeding the predetermined peak
30

expansion cylinder pressure.

Accordingly,

the intake valve

82 must be closed progressively either earlier or later ascompressor

strokeincreases

volumetricefficiency

in

to

order

sacrificetomaintaincompressor
expanderwo 2012/040431

PCT/US2011/052711

30

volumetric efficiency and to prevent the expansion cylinder
peak pressure from exceeding its limit.
BSFC and BMEP suffer.
length of 90 mm,
5For example,

As a result,both

at a compressor stroke

the BSFC has risen from 282 g/kw-hr to a

maximum value 305 g/kw-hr, and the BMEP has fallen from 12.8
bar to a minimum of 10 bar.
FIG.

6characteristics

illustrates

of agasolinethefueledvolumetricefficiency

split-cycle engine

while operating the Miller cycle at 4000 RPM and full10

conditions for various boost levels.

15

72)

which,

for ais typically exceeded before peak expandercylinder pressure
(66,

load

The limiting constraint

in this case was chosen to be knock fraction,
gasoline engine,

50

isexceeded.

and expander(68, 74)

As

before,thecompressorof engine 50 were set to havea fixed bore diameter of 87 mm, and the expander was also set
to have a fixed stroke length of 87 mm.

Accordingly,thedisplacement volume of the compressor was varied by changing
the size of the stroke length of the compression piston 72.
It
20

is

importantillustrated in FIG. 5,

to

note

that,

unlikethe intake valve closingof intake valve 82 was not varied.

Rather,

the

(IVC)

case
timingthe IVC timing

was fixed at its optimum trapped mass condition to maximize
volumetriccompressor

efficiency

at

0.90

or

greaterregardless of boost level or compressor displacement.
25

Referring to FIG.
expanderstrokevolumetriclengthefficiency

are

graph

6,

three pairs of compressor andefficiency
plotted.

132,

136,

graphs
Each

140

is

versuscompressor

compressor

volumetricreferenced

tointakemanifold conditions and each expander volumetric efficiency
30

graph

130,

134,

138

is

referenced

toambient

conditions.

Graphs 132 and 130 illustrate the volumetric efficiencies ofcompressor (66, 72) and expander (68, 74) respectively, for a
boost

level

of

1. 7 barabsolute,

overacompressor

strokewo 2012/040431

PCT/US2011/052711

31
range of 60 mm to 80 mm.
volumetric efficiencies
(68,

74)

respectively,

Graphs 136 and 134 illustrate theof compressor

(66,

72)

and expander

for a boost level of 2.0 bar absolute,

over a compressor stroke range of 55 mm to 70 mm.
5Graphs 140

and 138 illustrate the volumetric efficiencies of compressor
(66,

72)

and

expander(68,level of 2.3 bar absolute,74)

respectively,

for

a

boost

over a compressor stroke range of

45 mm to 65 mm.
In each of the above cases,
10

was

swept

until

limit

(142,

i.e.,

1.7,

the

identical

144 and 146)

wasthe compressor strokepredeterminedreached for

2.0 and 2.3 bar absolute.

15

ofapproximately

efficiency

of77

approximately

mm

and

of

approximately

efficiencyof68

1. 28.

approximately

mm

and

1. 37.

lengthofapproximately

62

mm

and

the

compressor stroke

expandervolumetricFor

2. 0 bar

boost,

thecompressor strokeanexpandervolumetric

For

2. 3

boost,

knock fraction limit 146 was reached at a20

each boost level,anknock fraction limit 144 was reached at alengthfraction

For 1.7 bar boost,knock fraction limit 142 was reached at alength

knock

an

bar

the

compressor stroke

expandervolumetricefficiency of approximately 1.43.
The

following

three

importantcharacteristics

ofgasoline split-cycle engine 50 can be derived from the familyof graphs
25

(130,

132,

134,

136,

138

and 140)

illustrated in

FIG. 6:
1)

When

trappedtheoperatingmass

intakeconditions,compressor

(132,

136

valve

closing

thevolumetric

and

140)

at

maximum

efficiency

remainsofessentiallyconstant regardless of boost level or compressor size.
30

2)

For an identical knock fraction limit

14 6),

as boost level increases

bar absolute)(142,

144 and(from 1. 7 to 2. 0 to 2. 3

the compressor size decreases respectively(from 77 mm to 68 mm to 62 mm) .

wo 2012/040431

PCT/US2011/052711

32

3)

For an identical knock fraction limit(142,

144, and146) the volumetric efficiency of the expander (130, 134
and

5

138)

increases(fromthoughcompressor

sizeboost.

Asaresult,efficiency

isincreasing,

1.28

is

must also increase and,

to

1.37

to

decreasing
since

even

with

increasing

expandervolumetric

the mass

therefore,

1.43)

air

and

fuel

flow

specific power also

increases.
Referring
10

to

FIG.

7,

a

temperature

versuscompression comparison illustrates another advantage of thepresent split-cycle engine utilizing Miller operation.
150

illustrates

withinthetemperaturecompression

versuscylindercompression

of

a

Graph

pressure

split-cycle

enginehaving a pure piston compression arrangement (i.e., naturally
15

aspirated,152

no Miller operation,and no turbocharger) .

Graphillustrates temperature versus compression pressure forthe compression cylinder of a

split-cycle engine having an

arrangement of turbocharging and intercooling prior to piston
compression
20

Milleroperation) .

Theturbocharged/intercooled graph 152 results in a significantly
lower

final

pressure40

bartemperaturecomparedgraph 150.

25

(i.e,

to

for

that

of

the

same

end

of

compression

the

pure

pistoncompression

For example, at an end of compression pressure ofabsolute,

the

for

turbocharged/intercooled

cylindertheshown in graph 152,

temperature

withinthecompression

arrangement,asis approximately 475 degrees centigrade,while the temperature within the compression cylinder of thepure piston compression arrangement,

as shown in graph 150,

is approximately 550 degrees centigrade.
30

This lower temperature advantage is generally true
for any turbocharged/intercooled Miller cycle engine, whether
it be conventional or split-cycle.
mentioned

split-cycle

engine's

However,abilitytothe previously
utilizemorewo 2012/040431

PCT/US2011/052711

33

aggressive Miller factor operation than conventional engines
allowstheoverallcompression

resulting5

thatsplit-cycle

of

in
a

a

engineprocesslower

tointercoolthanaMillerinconventional

end-of-compression

conventional

later

cycle

engine,temperatureengine

for

the

than

the

same

pressure.
FIGS.

8 - 10 illustrate the effect of turbocharger

efficiency on the performance gains of the Miller cycle inthe10

present

split-cycle

engine

50.

FIG.

8

depicts

representative values of overall turbocharger efficiency as afunction

of

boost

level

for

a

low

and

high

efficiency

turbocharger system at the two engine speeds of 1400 RPM and4000 RPM.
in
15

20

the

These turbocharger efficiency estimates were usedfollowingcycle

analysisengine.

More particularly,

overallturbocharger

graph

162

the

Millersplit-cycle

graph 160 represents optimistic

efficiency

representsofestimates

pessimistic

at

1400

overallRPM

andturbocharger

efficiency estimates at 1400 RPM.

Additionally,graph 164

represents

optimistic

turbocharger

efficiency

estimates at

4000

overall

RPM and graph 166 represents pessimistic

overall turbocharger efficiency estimates at 4000 RPM.
Although every

turbocharger

system isunique,itcan generally be said that overall turbocharger efficienciesof turbocharger systems at lower speeds
25(such as 1400 RPM) do

not diverge as much as the overall turbocharger efficienciesof thosesame turbocharger systems(such as 4000 RPM).

at higher engine

This is illustrated in FIG.

speeds8 wherein

the overall efficiencies at 1400 RPM stay relatively close invalue over the entire boost range of 1.7 bar boost to 2.9 bar
30

boost

for

both

the

high

graph

160

and

low

estimated

162.

For

example,estimatedtheefficiency

efficiency

turbocharger

turbocharger

efficiencies

vary

graph

betweenapproximately 47 and 53 for a 1.7 bar boost in graphs 162 andwo 2012/040431

PCT/US2011/052711

34

160

respectively,

and

the

efficiencies

vary

betweenapproximately 46 and 52 for a 2.9 bar boost in graphs 162 and160 respectively.
In contrast,5

the overall turbocharger efficiencies

at 4000 RPM have a pronounced divergence over the boost range
between the high estimated efficiency case graph 164 and thelow estimated efficiency case graph 166.

For example,

theefficiencies between 1.7 bar and 2.0 bar boost in graphs 166
and 164
10

boost,

are virtually identical.
the graphs

However,

16 6 and 164 diverge

beyond 2. 0 bar

substantially until,

when they reach 2.9 bar boost, the overall efficiency for the

15

high

estimated

164)

while

efficiency

theoverallturbocharger
efficiency

is

of

about

the

estimated

These

overallturbocharger

is

about

32.

turbocharger

efficiencies

and

their

respectiveaprofoundeffectonengine(graphlowefficiency

have

48

divergence

performanceinaturbocharged Miller split-cycle engine such as engine 50 asshown in FIGS. 9 and 10.
FIG. 9 depicts engine performance (in terms of BMEP
20

and BSFC)

at

over

same

the

low engine speed
range

of

(1400

boost

rpm)

levelsand at full

for

both

the

load
high

efficiency and the low efficiency turbochargers exemplified
in FIG.

25

8.

The compressor stroke length(i.e.,

compressorsize),

required in combination with the corresponding boost

level

to

maintainoverlaying

the

a

fixed

x-axis.

knock

fraction,

is

More

particularly,

also

shown

graph

170

illustrates BMEP for the high efficiency turbocharger at 1400

30

RPM and graph 172

illustrates BMEP

turbocharger

1400

at

RPM.

for

the

low efficiency

Additionally,

graph

174

illustrates BSFC for the high efficiency turbocharger at 1400
RPM and graph 17 6 illustrates BSFC

for

the

turbocharger at 1400 RPM.

As can be seen,turbocharger

atefficiencies

1400

RPM

low efficiency

since the overall
of

both

the

high

wo 2012/040431

PCT/US2011/052711

35
efficiency turbocharger and low efficiency turbocharger track
each

other

fairlyclosely

boost to 2.9 bar boost
of
5

FIG.

8) ,

so

too

10

10

the

do

the

BSFC

depicts

engineat high engine

load over thesame range of boost

efficiency and

the

FIG.

8.

The

curvesperformance

(in

terms

speed

(4000

levelsstrokeRPM)

BMEP

and atto

fixed

x-axis.

offull

for both the highlength

level

a

bar

and

required in combination with themaintain

1. 7

17 6

low efficiency turbochargers

compressor

the

of

17 4,

size),

overlaying

range

9.

BMEP and BSFC)

in

boost

(as illustrated in graphs 160 and 162

curves 170, 172 of FIG.
FIG.

over

exemplified(i.e.,

compressor

corresponding boost

knock

fraction,

is

More

particularly,

also

shown

graph

180

illustrates BMEP for the high efficiency turbocharger at 4000
15

RPM

and

graph

turbocharger

182

illustrates

at

4000

BMEP

RPM.

for

the

low

Additionally,

efficiency
graph

184

illustrates BSFC for the high efficiency turbocharger at 4000
RPM

and

graph

18 6

illustrates

BSFC

for

1.7

to

the

low

efficiency

barboost,

turbocharger at 4000 RPM.
20

Withinoveralltheturbocharger

range

of

efficiencies

at

2.0

4000

RPM

of

the

both

the

high efficiency turbocharger and low efficiency turbocharger
track each other fairly closely (as illustrated in graphs 164

25

and 166 of FIG.

8).

As a

result,the BSFC curves 184,

and

18 0,

182

of

9

BMEP

curves

FIG.

also

track

each

186

other

fairly closely over that same boost range of 1.7 to 2.0 bar.

30

However,

at higher boost levels of 2.0 bar to 2.9
turbocharger

bar,

the

overall

both

the

high

efficiency

turbocharger

turbocharger diverge rapidly
164

and

166

of

FIG.

8).

efficiencies
and

at

4000

low

RPM

of

efficiency(again as illustrated in graphs

Accordingly,

curves 184, 186 and BMEP curves 180,

theassociated BSFC

182 also diverge rapidly

over that same boost range of 2.0 to 2.9 bar.

wo 2012/040431

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36

A further comparison of FIGS. 9 and 10 reveals thatan acceptable
and

high

Based
5

on

tradeoff forengine
the

speed

fixed

compressor stroke length atoperation

expanderisapproximately

55

mm.

strokelength

of

87

mm,

the

about

63%

of

the

displacement volume

of

the

compressor

isdisplacement

of

the

expander(i.e.,

volume

low

55

mm/87

mm

sized

for

0. 63) .

Turning

to

FIG.

11,

engine50

isacompressor having a stroke length of 55 mm and an expander
10having a stroke length of 87 mm.

The bore diameters of bothcompressor and expander were fixed at 87 mm.
was varied from 1400 to 4000 rpm at fullEngine speed

load and a turbo

boost of 3 bar absolute or less was applied.

The overall

efficiency of the turbocharger system was assumed to be that15of the low efficiency turbocharger system of FIG.

8

(i.e.,

graphs 162 at 1400 RPM and graph 166 at 4000 RPM).
Accordingly in FIG. 11, graph 190 is the full load
BMEP for this embodiment of engine 50 over the entire speed
range of 1400 to 4000 RPM.
20

Graph 192 illustrates the peak

expansion cylinder pressure over the entire
bar-absolute(read from left y-axis).

speed range

inGraph 194 represents

the peak compression cylinder pressure over the entire speed
range in bar-absolute
resulting25

peak(also read from the left y-axis).

compression

cylinder66

pressure

and

The
peak

expansion cylinder 68 pressure, which generally increase withengine speed and BMEP,
factor,which

are low due to an aggressive Miller

providesincreased

BMEP

due

to

reducedcompression work instead of increased expansion work as most30

performance improvement schemes provide.

By way of example,

at

an

an

engine

speed

of

4000

RPM

and

engineBMEPofapproximately 19 bar, the peak expansion cylinder pressure isapproximately

94

bar-absolute

and

the

peak

compression

cylinder pressure is approximately 79 bar-absolute.

wo 2012/040431

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37

In general,
the

more

weight.

the lower the peak cylinder pressures,

advantageous

it

is

for

engineAccordingly,foragasoline

performance

split-cycle

andenginehaving an engine speed of 3500 RPM or greater and an engine
5

BMEP of 15 bar or greater,
expansion

cylinderpressure

desirable

to

the

less

110

than

expansion10

it is desirable to have the peak

have

bar,

Additionally,

for

most

gasoline

than

be

120

bar,

more

cylinder pressuredesirablepressurealesspeak expansion

andcylinder

be

to

less

split-cycle

have

be

the

peak

than

100

bar.

enginehavinganengine speed of 3500 RPM or greater and an engine BMEP of 15
bar or greater,

it is desirable to have the peak compression

cylinder pressure be

less

than

100

bar,

more

desirable

tohave the peak compression cylinder pressure be less than 90
15bar, and most desirable to have the peak compression cylinderpressure be less than 80 bar.
The invention and the embodiments of the invention
described herein are not fuel specific.

One skilled in theart would recognize that the invention may be utilized with
20

any

fuel

appropriate

combustion

enginefor

anyapplication

reciprocating(e.g.,

internal

gasoline,

diesel,natural gas or the like)
Althoughreference
25

to

the

inventionhas

specific embodiments,

it

been

described

byshould be understoodthat numerous changes may be made within the spirit and scopeoftheinventive

concepts

described.

Accordingly,itisintended that the invention not be limited to the describedembodiments,

but that it have the full scope defined by thelanguage of the following claims.

wo 2012/040431

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38
CLAIMS

What is claimed is:1.

A split-cycle engine comprising:a crankshaft rotatable about a crankshaft axis;

5

anvolume,expanderhavinganexpanderdisplacement

the expander including an expansion piston slidably

received

withinan

expansion

cylinderandoperatively

connected to the crankshaft such that the expansion piston
10

reciprocatesthroughanexpansion

strokeandanexhauststroke during a single rotation of the crankshaft;avolumecompressor

that

is

90

percent

displacement volume,15

havingaordisplacement

less

the

than

expanderthe compressor including a compression

piston slidably received within aoperatively

compressor

connected

tothecompression cylinder andcrankshaft

such

that

the

compression piston reciprocates through an intake stroke andacompression

strokeduringasinglerotationofthecrankshaft;20

acrossovercompression

cylindercrossover passage

passageandtheincluding at

interconnecting
expansion

least athecylinder,thecrossover expansion

valve disposed therein;an25

intakemanifoldconnected

tothecompression

cylinder;a boosting device connected to the intake manifold
and operable to provide a 1.7 bar absolute or greater boostpressure level to the intake manifold;30

manifoldanintakevalve

disposedandthecompression

betweentheintakecylinderandoperable

controlfluid communication therebetween,theintake valve

havinganintakevalve

closing

event

which

is

timed

to

to

wo 2012/040431

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39

provideacompressor

volumetricefficiency

relativetointake manifold conditions of 0.75 or greater; andthecompressor

displacement

volume

being

sized

relative to the expander displacement volume such that the5

combinationpressure

of

level

compressor

displacement

provideexpanderanvolumeandvolumetricboost

efficiency

relative to ambient conditions that is 0.90 or greater.
2.

The

split-cycle

engine

of

claim

1,

whereinthe compressor has a compressor displacement volume that is
10

80 percent or less than the expander displacement volume.
3.

The

split-cycle

engine

of

claim

1,

whereinthe compressor has a compressor displacement volume that is
75 percent or less than the expander displacement volume.
4.
15

The

split-cycle

engine

of

claim

1,

whereinthe compressor has a compressor displacement volume that is
70 percent or less than the expander displacement volume.
5.
theboosting

absolute
20

The

split-cycle

deviceisorgreater

boost

6.

The

engineoperabletopressure

of

claim

1,

whereinprovide

a

2.0

level

bar

to

the

intakeclaim

1,whereinprovidea

2.3

manifold.

theboosting

absolute

split-cycle

deviceisorgreater

boost

7.

The

engineoperabletopressure

of

level

bar

to

the

intakeclaim

1,whereinprovidea

2.5

manifold.
25
theboosting

split-cycle

deviceisorgreater

boost

8.

The

intakevalve

absolute

engineoperabletopressure

of

level

bar

to

the

intakeclaim

1,whereinmanifold.

30

thesplit-cycle
closing

engine

event

is

of

timed

to

provideacompressor volumetric efficiency relative to intake manifold
conditions of 0.80 or greater.

wo 2012/040431

PCT/US2011/052711

40

the

9.

The

intakevalve

split-cycle
closing

engine

event

is

of

claim 1,

timed

to

whereinprovideacompressor volumetric efficiency relative to intake manifold
conditions of 0.85 or greater.
5

10.
the

intakeThe
valve

split-cycle
closing

engine

event

is

of

claim 1,

timed

to

whereinprovideacompressor volumetric efficiency relative to intake manifold
conditions of 0.90 or greater.
11.
10

The

split-cycle

engineofclaim 1,

whereinthe compressor displacement volume is sized relative to the
expander displacement volume
compressor
providedisplacement

anexpandersuch thatvolumeandvolumetric

the

boost

combinationpressureefficiency

of

level

relativetoambient conditions that is 1.0 or greater.
15

12.

The

split-cycle

engineofclaim 1,

whereinthe compressor displacement volume is sized relative to the
expander displacement volume
compressor
provide20

displacement

anexpandersuch thatvolumeandvolumetric

the

boost

combinationpressureefficiency

of

level

relativetoambient conditions that is 1.1 or greater.
13.

The

split-cycle

engineofclaim 1,

whereinthe compressor displacement volume is sized relative to the
expander displacement volume
compressor
25

providedisplacement

anexpandersuch thatvolumevolumetric

and

the

boost

combinationpressureefficiency

of

level

relativetoambient conditions that is 1.2 or greater.
14.

The split-cycle engine of claim 1, wherein:

the boosting device is operable to provide a 2.0
bar absolute or greater boost pressure level to the intake
30

manifold;
the intake valve closing event is timed to provideacompressor

volumetricefficiency

relative

manifold conditions of 0.80 or greater; andtointakewo 2012/040431

PCT/US2011/052711

41

the

compressor

displacement

volumeissizedrelative to the expander displacement volume such that the
combinationpressure
5

of

level

compressor

displacement

provideexpanderanvolumeandvolumetricboost

efficiency

relative to ambient conditions that is 1.0 or greater.
15.

The split-cycle engine of claim 1, wherein:

the boosting device is operable to provide a 2.3
bar absolute or greater boost pressure level to the intake
manifold;10

the intake valve closing event is timed to provideacompressor

volumetricefficiency

relativetointakeissizedmanifold conditions of 0.85 or greater; andthecompressor

displacement

volumerelative to the expander displacement volume such that the15

combinationpressure

of

level

compressor

displacement

provideexpanderanvolumeandvolumetricboost

efficiency

relative to ambient conditions that is 1.1 or greater.
16.

The

split-cycle

engineofclaim 1,

whereinofclaim 1,

whereinthe boosting device is a turbocharger.
20

17.
the

The

compressor

split-cycle

engineincludesplurality

aofcompression

cylinders.
18.

The

split-cycle

engineofclaim 1,

whereinthe expander includes a plurality of expansion cylinders.
25

19.
a

crossoverpassage,

The split-cycle engine of claim 1,
compression

valve

disposedintheincludingcrossoverthe crossover compression valve and the crossover

expansion valve forming a pressure chamber therebetween.
20.
30

A method of operating a

split-cycle engine,the engine including:a crankshaft rotatable about a crankshaft axis;
anvolume,expanderhavinganexpanderdisplacement

the expander including an expansion piston slidably

wo 2012/040431

PCT/US2011/052711

42
received

withinan

expansion

cylinderandoperatively

connected to the crankshaft such that the expansion piston
reciprocatesthroughanexpansion

strokeandanexhauststroke during a single rotation of the crankshaft;5avolume,slidably

compressor

thecompressor

received

operatively

havingcompressor

including

withinconnected

a

to

a

displacement

compressionacompression

thecrankshaft

pistoncylinder
such

and

that

the

compression piston reciprocates through an intake stroke and10acompression

strokeduringasinglerotationofthecrankshaft;crossoveracompression

cylindercrossover passage
15

passageandtheincluding at

interconnecting
expansion

least athethecylinder,crossover expansion

valve disposed therein;anintakemanifoldconnected

tothecompression

cylinder;a boosting device connected to the intake manifold
and operable to provide a boost pressure level to the intake
20

manifold; andmanifoldanintakevalve

disposedandthecompression

betweencylinderandtheintakeoperabletocontrol fluid communication therebetween;
the method including the steps of:25

sizing the compressor displacement volume to be 90
percent or less than the expander displacement volume;deliveringpressure

level

to

a
the

1. 7

bar

intakeabsoluteorgreatermanifold

from

the

boost

boosting

device;30

timing the closing of the intake valve such that acompressor volumetric efficiency relative to intake manifold
conditions is 0.75 or greater; andwo 2012/040431

PCT/US2011/052711

43
sizing the compressor displacement volume relative
totheexpandercombinationofdisplacement

compressor

volumedisplacement

such

and

boost

that

the

pressurelevel provide an expander volumetric efficiency relative to5ambient conditions of 0.90 or greater.
21.

The

methodofclaim 20,

including

the

step

of:
sizing the compressor displacement volume to be 80
percent or less than the expander displacement volume.
10

22.

The

methodofclaim 2 0,

including

the

step

of:
sizing the compressor displacement volume to be 75
percent or less than the expander displacement volume.
23.
15

The

methodofclaim 20,

including

the

step

of:
sizing the compressor displacement volume to be 70
percent or less than the expander displacement volume.
24.

The

methodofclaim 20,

including

the

step

boost

of:
20

deliveringpressure

a

2. 0

level

to

the

25.

The

methodbar

intakeabsoluteorgreatermanifold

from

the

boosting

device.
ofclaim 2 0'

including

the

step

boost

of:
25

deliveringpressure

a

2.3

level

to

the

2 6.

The

methodbar

intakeabsoluteorgreatermanifold

from

the

boosting

device.
ofclaim 2 0'

including

the

step

boost

of:
delivering

30
pressure
device.

level

to

a
the

2.5

bar

intakeabsoluteorgreatermanifold

from

the

boosting

wo 2012/040431

PCT/US2011/052711

44
27.

The method ofclaim 20,

including

the

step

of:
timing the closing of the intake valve such thatthe5

compressor

volumetricefficiency

relativetointakemanifold conditions is 0.80 or greater.
28.

The method ofclaim 20,

including

the

step

of:
timing the closing of the intake valve such thatthe10

compressor

volumetricefficiency

relativetointakemanifold conditions is 0.85 or greater.
2 9.

The method ofclaim 2 0,

including

the

step

of:
timing the closing of the intake valve such thatthe15

compressor

volumetricefficiency

relativetointakemanifold conditions is 0.90 or greater.
30.

The method ofclaim 20,

including

the

step

of:
sizing the compressor displacement volume relative
to
20

the

expandercombinationofdisplacement

compressor

volumedisplacement

suchand boost

that

the

pressurelevel provide an expander volumetric efficiency relative toambient conditions of 1.0 or greater.
31.

The method ofclaim 20,

including

the

step

of:
25

sizing the compressor displacement volume relative
totheexpandercombinationofdisplacement

compressor

volumedisplacement

suchand boost

that

the

pressurelevel provide an expander volumetric efficiency relative toambient conditions of 1.1 or greater.
30

32.

The method ofclaim 20,

including

the

step

of:
sizing the compressor displacement volume relative
totheexpanderdisplacement

volume

such

that

the

wo 2012/040431

PCT/US2011/052711

45

combinationofcompressor

displacement

and boostpressurelevel provide an expander volumetric efficiency relative toambient conditions is 1.2 or greater.

wo 2012/040431

PCT/US2011/052711

1111

FIG. 1
Prior Art

wo 2012/040431

PCT/US2011/052711

2111

FIG. 2A

Miller Cycle with Late IVC
p3v

FIG. 28

Miller Cycle with Early IVC
p

3

2
4
Pi, Pe6 --------..;;;::::.,~--...... 57v

wo 2012/040431

PCT/US2011/052711

3/11

FIG. 3104

rH------l~Exha ust

108

Exhaust
110

112

98

50~

64

72

wo 2012/040431

PCT/US2011/052711

4111

FIG. 4
104

108
110

:.;;

FIG. 5

Miller Cycle: 1. 7 bar. limited PCP. variable IVC

0

120

N
0

......

I

1:' 310N0..c::: 305

~ 300

-

0
"""

~

"""
......
~,_looo""""-9 295___. ~290~ 285
en 2800~13,..--....

12

'-

co

..c

~ 11

w

~

aJ

122
10

~
-..l.
-..l.

9
1.20
1.10
1.00

II=

w

I

.g 0.90.........

...

.........

......

... ......
......

......

......

......

-- .. -------- ----------+-Con ~pressor(rei<: tive to intake mar ifold)--............. r----_~ ......

0.80

124

/

_ _ _ _ ... ___ -,jjl"d-&-Ex ander("")(rei l8tive to 8mbi1 nt)

~

0.70

-1 26

""""''

0.60

~
d

00
N
0

......
......

-0
Ut
N
-..l

......
......

0.50
50

55

60

70
75
65
Compressor Stroke (mm)

80

85

90

:.;;
0

FIG. 6N
0Miller Cycle Volumetric Efficiency Characteristics
4000 rpm Full Load......

N

0
0
"""

"""
......
~

1.50~1.7 boost, camp

146-r-... ;,')D-a- 2.( boost, camp..c::,_

1.40\,.../

.....

I) ':

h,...,...,..f """"''"'

vvvv•, vv''t-'-~-1. boost, exp--EJ-2. Dboost, exp

I-t"4-?

;: 1.30

'--'

a.>

"(3

~ 1.20

;

;

;

;

;~

,(

~
;

1.00

;

;

;

;

;

~'
;

D

;

;

;

;

;

;

"

~

50

...

;

;

;

... ...

... ...

~~~;

.....

;"

,,,

142 .. --< ~... ...
;

I...

t,. j..<lkii
ISI!f!l

I;?

~~

""'""

55

60

65

;

;

......

... .S:I>

IV knock

;

D

;

~
-..l.
-..l.

~'---130

(136

(140

45

~nock

;"'- f-'r--134r f•{<])' ...

'-

;

,""-

Lj-J

/"

;

rF-('", ..... h'I-'~;

0.90 '""

144-;
;

:>

--,

,.f:

;

b'

;

)

knoc

;

/"--138

(_)

1.10

;

;

;

;

·c
.......
a.>

;

;

(_)

::l

;

;

;

;

;

;

hnnc::t e:::wn

c

E

'"v""

,~

;

~~o.-.

1.:~

~

32

~

g

A

\..

70

75

Compressor Stroke (mm)

'"d

("")

80

00
N
0

......
......

0Ut

N
-.:t

Voi-Eff for comp. cylinder referenced to intake manifold.
Voi_Eff for exp. cylinder referenced to ambient......
......

:.;;

FIG. 7

0
N
0

"""

N

150

Compression Comparison0"'""'""""
0

~

600

--

500
152

0()
( ])

s....
::l

400

........

co
s....

(])

a..

E
~

300

~
-..l.
-..l.

200

- Piston compre~sion
Turbo comp--~piston camp

100
00

10

20

30

40

Pressure (bar-absolute)

50
'"d

("")

g
00

•

For same final compression pressure, after cooling part-way
through event reduces final temperature significantly

N
0

"""

"""
0Ut
N
-..l

"""
"""

:.;;

FIG. 80Turbocharger Efficiency Characteristics

N
0

70

......

N

0
0
"""

"""
......
~

16460-------50,..--....
I----------...!....-

>.
()

c

Q)

"(3

40

w

- - - 4000 rpm, hi h

~
-..l.

E->

--+-- 4000 rpm, lo

-..l.

tE

....
Q)

co
..c

()

30

0

..c
....
::::l

I-

co
....
Q)

>
0

20
'"d

("")

g

10

00
N
0

......
......

0Ut
N....:t......
......

0
1.7

2

2.3

2.6
Boost Level (bar-abs)

2.9

FIG. 9:.;;Miller Cycle 1400 RPM Performance

23

I

0
N
0

270

"""

N

BMEP, h gh turbo eff• ~BMEP, I w turbo eff-----BSFC, h gh turbo eff--EJ-BSFC, I wturboeff

0

•

21

--

fll"'

19

"'"
"'""""
0

~

260

250

-I....

..c

I....

~

co

-

..0
.._
0....

~

0}
.._

w

~
""""'
""""'

()
LL

~

C1J

240 (/)
C1J

17

15

y

'"d

1230

("")

g
00
N
0

"""

66.0

~

58.5
53.1
Correspfnding Compressor St1oke (mm)

48.9

2.3
2.6
Boost Level (bar-abs)

2.9

13
1.7

2

I 220

"""
0Ut
N
-..l

"""
"""

:.;;

FIG. 100Miller Cycle Knock-Limited Performance
4000 RPM

N
0
N
0

76.7

23

il

I

I

I

-

- ~- BMEP, low turbo eff

70.0
Ani"\

-

....
Q)

0

290

----+- BMIEP, high turbo eff
211 ~--""""'""'""""
~

I

186

280

I

BS C, high turbo eff,..--....

....
co,..--....

19..c::::::::

s

270 1::'
..c

63.3

~

---0....

---....

....

~

~

___ ..-,,_.. ........ ........

Q)w..,.,""".,s:---c~~----

.ttl'

-u
~

~

0

aJ

17

~
-..l.
-..l.

LL

0....

(f)

()

260 aJ

()
Q)

-..l.

56.7

0..
(f)

'"d

("")

15 1 - - - - ---~~---------+-----------+-----------~-----------1 250

g
00
N
0

"""

68.1
1

61.6

56.8

53.1

Corresfonding compressoriStroke (mm)'------43.3
2

2.3

2.6

Boost Level (bar-abs)

2.9

50.2

I 240

3.2

"""
0Ut
N
-..l

"""
"""

:.;;

FIG. 11

0
N
0

"""

Miller Cycle: 1400 to 4000 rpm: less than 3 bar absolute boost: full load:
BMEP and Cylinder Pressure Curves

co
.!...
co

......

v;i/"... ...

~--C190- ..... ------ -------lli80............

----- -- -- ----·IlL---+- Exp C I PCP

e

..-1

.......... camp Cyl PCP--k- BMEF

~

::::l
(/)
(/)
Q)

'-

0....

70

60

0

~

(192

20

~

U)

..c

"'"
"'""""

25

100

90

N

0

.....

~

~

15 1::'
co

e

K:

~

-..l.
-..l.

~
-..l.

0....

w

~

aJ

10

5

,/

'"d

("")

g
00
N
0

"""

"""
0Ut

50
1000

1500

2000

2500

3000

Engine Speed (rpm)

3500

4000

0

4500

N
-..l

"""
"""

INTERNATIONAL SEARCH REPORTInternational application No.
PCT/US 11/52711

A.

CLASSIFICATION OF SUBJECT MATTER

IPC(8)- F02B 19/00 (2012.01)
USPC- 123/253
According to International Patent Classification (IPC) or to both national classification and IPC
B.

FIELDS SEARCHED

Minimum documentation searched (classification system followed by classification symbols)
USPC: 123/253; F02B 19/00 (2012.01)

Documentation searched other than minimum documentation to the extent that such documents are included in the fields searched
123/253, ?OR, 261; 60/605.1; 606, 607; F02B 19/00, F02B$ (search term limited, see keywords below)

Electronic data base consulted during the international search (name of data base and, where practicable, search terms used)
PubWest (PGPB, USPT, EPAB, JPAB); Google (Patents, Scholar)
Search Terms; Engine, dynamo, turbine, power, train, Split-cycle, split, cycle, Crankshaft, shaft, Rotat$6, revolving, Axis, pivot, center,line, Expander, expansion, Displac$6, Exhaust, exit, volume, Piston, Stroke, cycle, Compress, compression, Crossover, Passage, Intake

C.

DOCUMENTS CONSIDERED TO BE RELEVANT
Citation of document, with indication, where appropriate, of the relevant passages

Category*

Relevant to claim No.

y

US 2009/0038597 A1 (Phillips) 12 February 2009 (12.02.2009), entire document, especially
Abstract, Fig 2, para [002]-[0027] and claim 11

1-32

y

US 2010/0116241 (Mistry) 13 May 2010 (13.05.201 0), entire document, especially abstract andpara [0002]-[0037]

1-32

y

GB 2,413,361 A (Maidment) 26 October 2005 (26.1 0.2005), entire document, especially pg 3,
para 52-4, 21-23yUS 7,513,224 B2 (Heaton) 07 April 2009 (07.04.2009), entire document, especially col 3, In 3343

17-18

y

US 6,952,923 B2 (Branyon et al.) 11 October 2005 (11.1 0.2005), entire document

1-32

y

US 2010/0095927 A1 (Salminen) 22 April2010 (22.04.0210), entire document

1-32

y

US 6,543,225 B2 (Scuderi) 8 April 2003 (08.04.2003), entire document

1-32

--

D

--

Further documents are listed in the continuation of Box C.

•

DSpecial categories of cited documents:'T""A'' document defining the general state of the art which is not considered
to be of particular relevance"E" earlier application or patent but published on or atler the international "X"
tiling date·'L" document which may throw doubts on priority claim(s) or which iscited to establish the publication date of another citation or other "Y"
special reason (as specified)"0'' document referring to an oral disclosure, use, exhibition or othermeans"P" document published prior to the international filing date but later than "&"the priority date claimed
Date of the actual completion of the international search
06 February 2012 (06.02.2012)
Name and mailing address of the ISA/US
Mail Stop PCT, Attn: ISAIUS, Commissioner for Patents
P.O. Box 1450, Alexandria, Virginia 22313-1450
Facsimile No. 571-273-3201

later document published after the international filing date or!eriority
date and not in conflict with the aphlication but cited to un erstandthe principle or theory underlying t e invention
document of particular relevance; the claimed invention cannot be
considered novel or cannot be considered to involve an inventive
step when the document is taken alone
document of particular relevance; the claimed invention cannot be
considered to involve an inventive step when the document iscombined with one or more other such documents, such combination
being obvious to a person skilled in the art
document member of the same patent family

Date of mailing of the international search report

14 FEB 2012
Authorized officer:
Lee W. YoungPCT Hclpdcsk: 571-272-4300
PCT OSP: 571-272-7774