111111

c12)

US008006656B2

United States Patent(10) Patent No.:

Branyon et al.

(45) Date of Patent:(54)

SPLIT-CYCLE FOUR-STROKE ENGINE(75)

Inventors: David P. Branyon, San Antonio, TX(US); Jeremy D. Eubanks, San Antonio,
TX (US)(73)

Assignee: Scuderi Group, LLC, West Springfield,
MA(US)( *)Notice:(21)1111111111111111111111111111111111111111111111111111111111111

Aug. 30, 2011

123/51 BD; 60/597, 653, 668, 698, 706,
712, 727, 729
See application file for complete search history.
(56)

References Cited
U.S. PATENT DOCUMENTS
810,347 A
848,029 A
939,376 A

Subject to any disclaimer, the term of thispatent is extended or adjusted under 35
U.S.C. 154(b) by 50 days.

111906 Porter
3/1907 Haselwander
1111909 Appleton(Continued)
FOREIGN PATENT DOCUMENTS
2515271DE

Appl. No.: 12/426,230

US 8,006,656 B2

10/1976(Continued)
(22)
(65)

Filed:

Apr. 18, 2009

OTHER PUBLICATIONS

Prior Publication Data
US 2009/0199829 A1

Aug. 13, 2009

Notice of Opposition against EP 1,925,795 B1 filed by Bernard
Golibrodski.

(Continued)
Related U.S. Application Data
(63)

Continuation of application No. 12/322,676, filed on
Feb. 5, 2009, now Pat. No. 7,810,459, which is acontinuation of application No. 12/283,522, filed on
Sep. 12, 2008, which is a continuation of application
No. 11/890,360, filed on Aug. 6, 2007, now
abandoned, which is a continuation of application No.
11/197,999, filed on Aug. 4, 2005, now Pat. No.
7,588,001, which is a continuation of application No.
10/864,748, filed on Jun. 9, 2004, now Pat. No.
6,952,923.

(60)

Provisional application No. 60/480,342, filed on Jun.
20, 2003.

(51)

Int. Cl.
F02B 25100(2006.01)
U.S. Cl. ........................ 123/70 R; 123/52.2; 123/68
Field of Classification Search ................. 123/52.1,
123/52.2, 52.3, 431, 53.1, 53.3, 53.5, 55.4,
123/55.5, 55.2, 68, 70 R, 51 R, 51 AC, 51 BC,

(52)
(58)

Primary Examiner- Michael Cuff
Assistant Examiner- Hung Q Nguyen(74) Attorney, Agent, or Firm- Fildes & Outland, P.C.

(57)
ABSTRACTAn engine has a crankshaft, rotating about a crankshaft axis of
the engine. An expansion piston is slidably received within an
expansion cylinder and operatively connected to the crankshaft such that the expansion piston reciprocates through an
expansion stroke and an exhaust stroke of a four stroke cycle
during a single rotation of the crankshaft. A compression
piston is slidably received within a compression cylinder andoperatively connected to the crankshaft such that the compression piston reciprocates through an intake stroke and acompression stroke of the same four stroke cycle during thesame rotation of the crankshaft. A ratio of cylinder volumes
from BDC to TDC for either one of the expansion cylinderand compression cylinder is fixed at substantially 26 to 1 orgreater.
10 Claims, 30 Drawing Sheets

162

154

US 8,006,656 B2
Page 2

U.S. PATENT DOCUMENTS
1,062,999 A
1,111,841 A
1,248,250 A
1,301,141 A
1,372,216 A
1,392,359 A
1,751,385 A
1,856,048 A
1,904,816 A
1,969,815 A
2,091,410 A
2,091,411 A
2,091,412 A
2,091,413 A
2,154,856 A
2,269,948 A *2,280,712 A
2,769,435 A *2,957,455 A
2,974,541 A
3,623,463 A *3,774,581 A
3,895,614 A
4,170,970 A *4,215,659 A
4,344,405 A *4,450,754 A
4,506,634 A
4,628,876 A
4,696,158 A
4,783,566 A
4,805,571 A
4,945,866 A
4,955,328 A
5,146,884 A
5,158,047 A
5,499,605 A
5,546,897 A
5,623,894 A
5,711,267 A
5,799,636 A
5,857,436 A *5,950,579 A
5,964,087 A
5,992,356 A
6,058,901 A
6,202,416 B1
6,230,671 B1
6,340,004 B1
6,415,749 B1
6,543,225 B2
6,609,371 B2
6,655,327 B1
6,722,127 B2
6,789,514 B2
6,817,185 B2
6,874,453 B2
6,880,502 B2
6,952,923 B2
6,986,329 B2
7,007,639 B1
7,017,536 B2
7,121,236 B2
7,353,786 B2
7,481,190 B2
7,588,001 B2

5/1913
9/19141111917
4/19193/1921
10/1921
3/19304/1932
4/1933
8/1934
8/1937
8/1937
8/1937
8/1937
4/1939111942
4/19421111956
10/1960
3/196111119711111973
7/1975
10/1979
8/19808/1982
5/1984
3/1985
12/1986
9/19871111988
2/19898/19909/19909/1992
10/1992
3/1996
8/1996
4/1997111998
9/19981119999/199910/19991111999
5/20003/20015/20011120027/20024/20038/2003
12/20034/20049/20041112004
4/20054/2005
10/2005
1120063/20063/2006
10/2006
4/20081120099/2009Webb
KoenigBohlerLeadbetter
Casaday
RudvistBeaudry
AhrensBeaudry
Meyer
Mallory
Mallory
Mallory
Mallory
Mallory
Mallory
Mallory
Cass
Bonvy
Dolza
DeVries .
Lundv
BaileyMcCandless
Lowther
ZaharisLiljequist
KerrebrockFujikawa
DeFrancisco
Kocal eta!.
Humphrey
ChabotSobotowski
MerkelSchaalThringBrackett
ClarkeWilliams
Fish
Chen.
Ott
Tort-Oropeza
Howell-Smith
Lee
Gray
Achterberg
PattonSturman et a!.
Scuderi
Scuderi
HedmanScuderi et a!.
SuhConey eta!.
Coney eta!.
Scuderi
Branyon et a!.
Scuderi et a!.
Luttgeharm
Scuderi
Scuderi et a!.
Scuderi et a!.
Scuderi
Branyon et a!.

FOREIGN PATENT DOCUMENTS
DE
EP

2628155
1290326 B1

111978
10/2005

123/52.3
123/68
123/70 R
123/52.4
123/68

FR
GB
GB
GB
IT
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP
JP

su

wo
wo

24163442996023838667210255055768-503043
51-39306
51-91416
54-89108
56-8815
56-99018
56-145641
57-181923
60-143116
60-256642
62-126523
63-124830
5-502707
5-156954
6-1598368-158887
8-232675
8-261004
2000-508403
2001-12250
2001-207801
2002-506949
1551880 A1
0116470 A1
03/046347

8/197911119281111932
12/1954
12/1954
4/1976
4/1976
8/1976
7/19792/1981111198111119811111982
9/1985
12/1985
8/1987
5/1988
5/1993
6/1993
6/1994
6/1996
9/1996
10/1996
7/20001/20018/20013/2002
3/19903/20016/2003OTHER PUBLICATIONS

123/70 R

English Translation ofNotice of Opposition against EP 1,925,795 B1
filed by Bernard Golibrodski.
English Translation ofEP 1290326 B1 to Bernard Golibrodski.
Tiainen eta!., Novel Two-Stroke Engine Concept, Feasibility Study,
SAE Technical Paper Series 2003-01-3211, Oct. 27-30, 2003.
Gronlund eta!., Valve Train Design for aNew Gas Exchange Process,
SAE Technical Paper Series 2004-01-0607, Mar. 8-11, 2004.
Search Report by Patro Information K.K. ofTokyo, Japan, dated Jun.
20, 2002, signed by Teruaki Haga, Managing Director.
Thermodynamic Modeling of the Two-Cylinder Regenerative Internal Combustion Engine by Francisco Ruiz, SAE 1991 Transactions,
Journal of Engines, Sec 3, vol. 100, p. 512-529.
Standard Search Report, No. RS 108446 US by the European Patent
Office, dated May 23, 2003.
U.S. Appl. No. 05/955,896, filed Oct. 30, 1978, titled "Internal Combustion Engine".
U.S. Appl. No. 05/955,895, filed Oct. 30, 1978, titled "Two Stroke
Internal Combustion Engine".
U.S. Appl. No. 05/947,998, filed Oct. 2, 1978, titled "Hybrid
Engine".
M.W. Coney, C. Linnemann, K. Sugiura and T. Goto, "First Prototypeof the High-Efficiency Isoengine", Institution of Diesel and Gas
Turbine Engineers (IDGTE), Power Engineer Journal, vol. 7, No.2,
Paper539, Sep.2004.
M.W. Coney, C. Linnemann, R.E Morgan, T.G. Bancroft and R.M
Sammut, "A Novel Internal Combustion Engine with Simul-taneous
Injection of Fuel and Pre-Compressed Pre-Heated Air", Proc. Fall
Technical Conference of the ASME Internal Combustion Engine
Division, Sep. 8-11, 2002, New Orleans, USA, ICE-vol. 39, pp.
67-77.
M.W. Coney, A. Cross, C. Linnemann, R.E. Morgan and B. Wilson,"Engineering Aspects of a Novel High Efficiency Reciprocating
Internal Combustion Engine", Proc. International Joint Power Generation Conference (ASME IJPGC 02), Jun. 24-26, 2002, Scottsdale,
Arizona, USA.

* cited by examiner

~

00
•

FIG.l
PRIOR ART~

FIG. 2PRIOR ART

~

~

~

=
~

~

~

(.H

~0

N
0

........

rFJ

=....

('D
('D

.....

0

.....

(.H

0

INTAKE

10COMPRESSION

rJl
00

10
~42

d

=
=
0'1

0..,

u.

0'1

=
N

~

00
•

FIG. 3
PRIOR ART~

FIG. 4PRIOR ART

~

~

~

=
~

~

~

(.H

~0

N
0

........

rFJ

=-

('D
('D

.....

N
0

.....

(.H

0

12
POWER STROKEAT MAX. PRESSURE
54

EXHAUST

d

10
10

rJl
00

=
=
0'1

0..,

u.

0'1

=
N

U.S. PatentAug. 30,2011

FIG. 5PRIORART

78

72___../
73___....,.._

70- ,_

Sheet 3 of 30

US 8,006,656 B2

80

82 ,.....-----'-----, 84-74
75

U.S. PatentAug. 30, 2011

Sheet 4 of 30

US 8,006,656 B2

FIG. 6

J~~l:O~~IB~fffll~~~li~

116-

128

100~

U.S. PatentAug. 30, 2011

Sheet 5 of 30

FIG. 7
152~

.

.....

~

. ..

US 8,006,656 B2

U.S. PatentAug. 30, 2011

US 8,006,656 B2

Sheet 6 of 30

FIG. 8
150

154

'0

~::::~iiiiin--1~..,+--]14

IIJ.,L..L..-,;-,.......~J..j-1-+~-+-r-138

U.S. PatentAug. 30, 2011

FIG. 9152-31o-1

Sheet 7 of 30

US 8,006,656 B2

162

170

182
~-114
ILIJ.,I.....-..,.,"""""""J.J..

140 _......,.___"138

U.S. PatentAug. 30, 2011

US 8,006,656 B2

Sheet 8 of 30

FIG. 9A

162

152-----~

182
116
106
140

'
/ '
/

'

'
/

/
/

f

/
/

/

/

/
/

/

/

/
/

114
138

/

'/
/

/

/

/
/

/

104

/

108

U.S. PatentAug. 30, 2011

Sheet 9 of 30

US 8,006,656 B2

FIG.JO
150

138

U.S. PatentAug. 30, 2011

Sheet 10 of 30

FIG. II150

US 8,006,656 B2

~

00
•
~

~

~

~

=

FIG.12A

~

218
~

§

I

227.:..,1.;/

L..,__J~r------1

217

-

vil-1

-

I ~

-

vel-1

>
=

L..,__J

229

~

236

cyll

~

233

(.H

~0

N
0

........

240
rFJ=........

('D
('D

~ I

~

ve2-1

-

L..,__J
r------1

238

I ~ I

.....

0

.....

(.H

0

241

[Q
INTAKE~

CrankTrai~
Exhaust

d

rJl
00

=
=
0'1

0..,

u.

0'1

=
N

U.S. PatentAug. 30, 2011

Sheet 12 of 30

US 8,006,656 B2

FIG. 12B
ITEM DEFINITION: CONVENTIONAL ENGINE• 211,212 intake manifold junctions/tees• intake: intake end environment (infinite
ambient source)• 214,215,216,217: intake ports• 218,219,220,221 fuel injectors• vix-y: intake valves• cyl1: conventional engine cylinder #1• cyl2: conventional engine cylinder #2• vex-y exhaust valves• 229,230,231,232 exhaust ports• 233,234: exhaust junctions or tees
• exhaust: exhaust end environment (infinite
ambient dump)• cranktrain: mathematical item for summing
items from all engine cylinders and handlingorganization such as firing order• 224-232 (not 228): "orifice" connections tohandle connections between pipes andjunctions in the model. These Items do not
represent anything in actual hardware.
• 243,248: mathematical links representing
the mechanical connection between
cylinders and the cranktrain.

~

00
•
~

FIG.13~~TOTAL COMBUSTION EVENT

1.0

~

~

~

=
~

/v
0.8

50o/o
BURN
POINT/
0.6

/

0.4

0.2
~

0.0
0

v
v

20

~

~

~

/

(.H

~0

N
0

........

/

rFJ

=....

('D
('D

.....

v
/

(.H

0

.....

(.H

0

10-90%COMBUSTIONd

DURATION~

rJl
00

I

40

60

CRANK ANGLE (deg)

80

100

=
=
0'1

0..,

u.

0'1

=
N

~

00

•

... Brake Torque (ft-lb)• Brake Power (hp)• Vol.Eff(%)• BTE(%)* FMEP (psi)FIG.14
100

~

~

~

~

=
~

~

~

90

(.H

~

-------.

....

80
70

60
50

~0

N
0

........

rFJ

40

30
20
10
0

1000

- --- ----=----------------

=....

('D
('D

.....

___.

.j;o.

0

.....

(.H

0

......

....

I

I

I

I

1500

2000

2500

3000

d

3500

Engine Speed (rpm)Summarv of Conventional En!!ine Performance versus En!!ine Soeed

rJl
00

=
=
0'1

0..,

u.

0'1

=
N

~

00

•

~

~

~

~

=

FIG.15A

§

I

~

227

L-...J
r------1

Inport lvil-2

-

>
=

~

~
I Comp

(.H

~0

N
0

........

exhportl

~

Power

ve2

rFJ
L-...J

245~

FYh-ir.n

=....

('D
('D

.....

Ul

0

.....

(.H

0

247

[Q
Intake~Canktrain~

Exhaust

d

rJl
00

=
=
0'1

0..,

u.

0'1

=
N

U.S. PatentAug. 30, 2011

Sheet 16 of 30

US 8,006,656 B2

FIG.15B
ITEM DEFINITION: SPLIT-CYCLE ENGINE• intake: intake end environment(infinite ambient source)• intk-bypass: single intake port before
split• intk-s_Plitter: intake manifold
junction/tee• intport 1, intport 2: intake ports• fuelinj 1, fuelinj 2, fuel injectors• vil-y: intake valves• comp: compression cylinder• check: check or reed valve atdischarge of compression cylinder• crossover 1: crossover passage• cross-valve: actuated crossover valve• power: expansion or power cylinder• vex: exhaust valves• exhport 1, exhport 2: exhaust ports• exh-jcn: exhaust junction/tee• exhaust: exhaust end environment (infinite ambient dump)• cranktrain: mathematical item for summing items from allengine cylinders and handling organization such as firing order.
• 222,223,226,227,245,246,247: "orifice" connections to handle
connections between pipes and junctions in the model.
These items do not representanything in actual hardware.

• 243,248 mathematical links representing the mechanicalconnection between cylinders and the cranktrain.

U.S. PatentAug. 30, 2011

Sheet 17 of 30

US 8,006,656 B2

~

00
•
~

~

~

FIG.17~--POWER PISTJON- - - - COMPRESSIONPJSTJON
TDC-

0

=
~

~

~

(.H

~0

N
0

1

........

2. xvc

xvo

rFJ

=....

('D
('D

PJSTION DISTANCE 3
FROM TDC (inches)
4 I

.....

QO

1':0

I \E.VO '~-- __. -;-;:~ A

BV~I

0

.....

(.H

0

5

6 ~------~~---~------~--------~---------~
0
180
270
90
360

d

rJl
00

=
=
0'1

CRANK ANGLE (deg ATDC-POWER CYL)

0..,

u.

0'1

=
N

~

00
•

FIG.18100
90• Indicated Torque (ft-lb)• Indicated Power (hp)
• Net IMEP (psi).

-------._______

80

~

~

~

~

=
~

~

~

---.

(.H

~0

N
0

........

70
60

~

rFJ

50

=....

:::::::.

40

30
20
10
0

1000

('D
('D

.....

...--- ~

\0

0

.....

(.H

0

I

I

I

I

1500

2000

2500

3000

d

3500

Engine Speed (rpm)Summsary of Initial Split-Cycle Engine Performance versus Engine Speed

rJl
00

=
=
0'1

0..,

u.

0'1

=
N

~

00
•
~

~

~

~

=

FIG. 19 Log P-LogV Diagram~

2 EngCylinder part cy11
10 __~~---------~

10-90°/o Burn=24 CA

~

(.H

~0

N
0

........

Expansion

9

8

rFJ

Pressure (Bar) ~

=-

('D
('D

.....

N
0
0

Compression

.....

(.H

0

Beathing Loop
d

2

3

4
5
VolumeNmakOx

6

7

8 9

rJl
00

=
=
0'1

0..,

Log-Log Pessure-Volume Diagam for Conventional Engine

u.

0'1

=
N

~

00
•

FIG.20~~~~P-V Diagram
EngCylinder part Power62.5.------------------------,Start of50.0 1-\ /Combustion

=
~

~

~

(.H

~0

10-90o/o Burn==16 CA

N0........

37.5
rFJ

Pressure (Bar)

=....

('D
('D

.....

N

25.0

0

.....

(.H

0

12.5
0.0 L..:====::r:::::====:::::r::======r::::::====::::I:::====:::::::1
0.008
0.208
0.408
0.608
0.808
1.009
Volume/V max
Pressure-Volume Diagram for Split-Cycle Engine

d

rJl
00

=
=
0'1

0..,

u.

0'1

=
N

~

00

•

• Conventional Engine• Intial Split-Cycle• 30-mm Crossoverx 20 TDC Phasing* 16 10-90% Bum• Crossover Ceramic Coating+ Expansion Cyl Cermnics

FIG. 21
45
44
43
42
41
40

ITE (%)

3
3
37
3

.....
-..

I

3
34
33
32
31
30
1000

~

~

=
~

~

~

-...________--.

--

...

~

~

(.H

~0

N
0

........

~

~

---.

.....

-

::::::--

~

rFJ

=-

('D
('D

.....

N
N
0

.....

~

(.H

0

~

~

d

I

1500

I

I

2000
2500
Engine Speed (rpm)

I

3000

3500

rJl
00

=
=
0'1

0..,

u.

Comparison of Split-Cycle Indicated Thermal Efficiencies

0'1

=
N

U.S. PatentAug. 30, 2011

Sheet 23 of 30

US 8,006,656 B2

FIG. 22258. J::
35% BURN OVERLAP

252

U.S. PatentAug. 30, 2011

US 8,006,656 B2

Sheet 24 of 30

FIG. 23256~-+---250

258

252

5% BURN OVERLAP

~

00
•
~

~

FIG. 24

~

~

=
~

0.006

~

0.005

~

(.H

~0

0.004

N
0

........

0.003
NOx(g)

0.002

{
I
I

0.001

-

30

-o~Jo1

I

--

10

50

70

N
Ul

.....

0

.-' '

30

=-

.....

(.H

,

-,

rFJ
('D
('D

0

--------------

I

I
./

-5o

- - Conventional---- Split-Early (35% Burn ov.erlap)------ Split-Late ( 5% Burn overlap)--------/

90d

Crank Angle (degrees)
Predicted NO emissions for all thee cases run
X

rJl
00

=
=
0'1

0..,

u.

0'1

=
N

~

00

•

~

~

~

FJG.25

~

=
~

3000

>
=
~

'-..

'-..

,,
'-..

(.H

"-..,

0

'-..
"-..,

~

'-..

'-- -....._

Expansion Piston
Thrust Load (N).,..... ___

90 105 .120 135 150 1651180

N0........

rFJ

=-

('D

a

N
0\
0

--Split-C ycle 10° l0-90(Yo Burn--- Split-Cycle 2i 10-90% Burn
-4000

.....

(.H

0

d

Crank Angle ( CA)

rJl
00

=
=
0'1

Expansion Piston Thrust Load as a Function of Bun1 Duration

0..,

u.

0'1

=
N

~

00
•

•
•

FIG.26

Indicated Power (hp)
ITE (0/o)

~

~

~

~

=
~

40
~

35

~

(.H

~0

N
0

........

30
25

rFJ

=-

('D
('D

.....

20

N
-....l
0

.....

(.H

0

15
10

0

20

40

60

80

100

120

Compression Ratio (Compression Cylinder)
Indicated Power and Thermal Efficiency versus Compression Ratio

d

rJl
00

=
=
0'1

0..,

u.

0'1

=
N

~

00

•

~

~

Power (hp)•• IndicatedITE (%)

FIG.27
40
38
36
34
32

~

=
~

~

--

~

-/

~

~

(.H

~0

N
0

........

rFJ

=-

30
28

('D
('D

.....

N
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....

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80
100
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Expansion Ratio (Expansion Cylinder)

120

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=
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Indicated Power and Thermal Efficiency versus Expansion Ratio

0..,

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Power (hp)•• IndicatedITE (%)

FIG.28
40
38
36
34

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28

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26
24

....

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12

14

16

18

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TDC Phasing (CA)

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Indicated Power (hp)
ITE (o/o)•-- -------.__

40
38
36
34
32

=
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30

N
0

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28
26

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-----------._

24

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0
0

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--------------.

22

20

~

25

30

35

40

45

50

55

60

65

70

Crossover Valve Duration ( CA)
Indicted Power and Thennal Efficiency versus Crossover Valve
Duration and Lift

(.H

0

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US 8,006,656 B2
1
SPLIT-CYCLE FOUR-STROKE ENGINE

2

strokes of one of more pistons for every expansion (or power)
stroke, i.e. for every stroke that delivers power to a crankshaft.
Referring to FIGS. 1-4, an exemplary embodiment of a
CROSS REFERENCE TO RELATED
prior art conventional four stroke cycle internal combustion
APPLICATIONS
5 engine is shown at 10. The engine 10 includes an engine block
12 having the cylinder 14 extending therethrough. The cylinThis patent application is a continuation application of
U.S. application Ser. No. 12/322,676, filed Feb. 5, 2009, titledder 14 is sized to receive the reciprocating piston 16 therein.
Split-Cycle Four-Stroke Engine, which is a continuation
Attached to the top of the cylinder 14 is the cylinder head 18,
applicationofU.S. application Ser. No. 12/283,522, filed Sep.
which includes an inlet valve 20 and an outlet valve 22. The
1012, 2008, titled Split-Cycle Four-Stroke Engine, which is abottom of the cylinder head 18, cylinder 14 and top (or crown24) of the piston 16 form a combustion chamber 26. On thecontinuation application ofU.S. application Ser. No. 11/890,
360, filed Aug. 6, 2007, titled Split-Cycle Four-Stroke
inlet stroke (FIG. 1), a air/fuel mixture is introduced into theEngine, which is a continuation application of U.S. applicacombustion chamber 26 through an intake passage 28 and thetion Ser. No. 11/197,999, filedAug. 4, 2005, titled Split-Cycle 15 inlet valve 20, wherein the mixture is ignited via spark plug
30. The products of combustion are later exhausted throughFour-Stroke Engine, which is a continuation application ofoutlet valve 22 and outlet passage 32 on the exhaust stroke
U.S. application Ser. No. 10/864,748, filed Jun. 9, 2004, now(FIG. 4). A connecting rod 34 is pivotally attached at its top
U.S. Pat. No. 6,952,923, titled Split -Cycle Four-Stroke
distal end 36 to the piston 16. A crankshaft 38 includes aEngine, which claims the benefit ofU.S. provisional application Ser. No. 60/480,342, filed on Jun. 20, 2003, titled Split- 20 mechanical offset portion called the crankshaft throw 40,
which is pivotally attached to the bottom distal end 42 of
Cycle Four-Stroke Engine, all of which are herein incorporated by reference in their entirety.
connecting rod 34. The mechanical linkage of the connecting
rod 34 to the piston 16 and crankshaft throw 40 serves to
FIELD OF THE INVENTION
convert the reciprocating motion (as indicated by arrow 44) of
25 the piston 16 to the rotary motion (as indicated by arrow 46)
of the crankshaft 38. The crankshaft 38 is mechanically
The present invention relates to internal combustion
linked (not shown) to an inlet camshaft 48 and an outlet
engines. More specifically, the present invention relates to acamshaft 50, which precisely control the opening and closing
split-cycle engine having a pair of pistons in which one pistonof the inlet valve 20 and outlet valve 22 respectively. The
is used for the intake and compression stokes and anotherpiston is used for the expansion (or power) and exhaust 30 cylinder 14 has a centerline (piston-cylinder axis) 52, whichis also the centerline of reciprocation of the piston 16. The
strokes, with each of the four strokes being completed in one
crankshaft 38 has a center of rotation (crankshaft axis) 54.
revolution of the crankshaft.
Referring to FIG. 1, with the inlet valve 20 open, the piston
16 first descends (as indicated by the direction of arrow 44) on
BACKGROUND OF THE INVENTION
35 the intake stroke. A predetermined mass of a flmable mixture of fuel (e.g., gasoline vapor) and air is drawn into theInternal combustion engines are any of a group of devices
in which the reactants of combustion, e.g., oxidizer and fuel,combustion chamber 26 by the partial vacuum thus created.
and the products of combustion serve as the working fluids of
The piston continues to descend until it reaches its bottom
dead center (BDC), i.e., the point at which the piston isthe engine. The basic components of an internal combustion
engine are well known in the art and include the engine block, 40 farthest from the cylinder head 18.
Referring to FIG. 2, with both the inlet 20 and outlet 22
cylinder head, cylinders, pistons, valves, crankshaft and camvalves closed, the mixture is compressed as the piston 16
shaft. The cylinder heads, cylinders and tops of the pistonstypically form combustion chambers into which fuel andascends (as indicated by the direction of arrow 44) on theoxidizer (e.g., air) is introduced and combustion takes place.
compression stroke. As the end of the stroke approaches top
Such an engine gains its energy from the heat released during 45 dead center (TDC), i.e., the point at which the piston 16 is
the combustion of the non-reacted working fluids, e.g., theclosest to the cylinder head 18, the volume of the mixture isoxidizer-fuel mixture. This process occurs within the engine
compressed in this embodiment to one eighth of its initial
volume (due to an 8 to 1 Compression Ratio). As the piston
and is part of the thermodynamic cycle of the device. In all
internal combustion engines, useful work is generated fromapproaches TDC, an electric spark is generated across thethe hot, gaseous products of combustion acting directly on 50 spark plug (30) gap, which initiates combustion.
moving surfaces of the engine, such as the top or crown of aReferring to FIG. 3, the power stroke follows with both
piston. Generally, reciprocating motion of the pistons is transvalves 20 and 22 still closed. The piston 16 is driven downferred to rotary motion of a crankshaft via connecting rods.
ward (as indicated by arrow 44) toward bottom dead center
Internal combustion (IC) engines can be categorized into(BDC), due to the expansion of the burning gasses pressing on
spark ignition (SI) and compression ignition (CI) engines. SI 55 the crown 24 of the piston 16. The beginning of combustion inconventional engine 10 generally occurs slightly before pisengines, i.e. typical gasoline engines, use a spark to ignite theton 16 reaches TDC in order to enhance efficiency. When
air/fuel mixture, while the heat of compression ignites the
air/fuel mixture in CI engines, i.e., typically diesel engines.
piston 16 reaches TDC, there is a significant clearance volThe most common internal-combustion engine is the fourume 60 between the bottom of the cylinder head 18 and thestroke cycle engine, a conception whose basic design has not 60 crown 24 of the piston 16.
Referring to FIG. 4, during the exhaust stroke, the ascendchanged for more than 100 years old. This is because of its
simplicity and outstanding performance as a prime mover ining piston 16 forces the spent products of combustion throughthe open outlet (or exhaust) valve 22. The cycle then repeats
the ground transportation and other industries. In a fourstroke cycle engine, power is recovered from the combustion
itself. For this prior art four stroke cycle engine 10, four
process in four separate piston movements (strokes) of a 65 strokes of each piston 16, i.e. inlet, compression, expansion
single piston. Accordingly, a four stroke cycle engine isand exhaust, and two revolutions of the crankshaft 38 arerequired to complete a cycle, i.e. to provide one power stroke.
defined herein to be an engine which requires four completeUS 8,006,656 B2
3

4

Problematically, the overall thermodynamic efficiency of
having a crankshaft, rotating about a crankshaft axis of theengine. An expansion piston is slidably received within an
the typical four stroke engine 10 is only about one third (1;3 ).
expansion cylinder and operatively connected to the crankThat is, roughly 1/3 of the fuel energy is delivered to the
crankshaft as useful work, lj3 is lost in waste heat, and 1/3 isshaft such that the expansion piston reciprocates through anlost out of the exhaust. Moreover, with stringent requirements
expansion stroke and an exhaust stroke of a four stroke cycleon emissions and the market and legislated need for increased
during a single rotation of the crankshaft. A compression
efficiency, engine manufacturers may consider lean-burn
piston is slidably received within a compression cylinder andtechnology as a path to increased efficiency. However, as
operatively connected to the crankshaft such that the comlean-burn is not compatible with the three-way catalyst, thepression piston reciprocates through an intake stroke and a
increased NOx emissions from such an approach must be 10 compression stroke of the same four stroke cycle during thedealt with in some other way.
same rotation of the crankshaft. A ratio of cylinder volumes
Referring to FIG. 5, an alternative to the above describedfrom BDC to TDC for either one of the expansion cylinder
conventional four stroke engine is a split-cycle four stroke
and compression cylinder is substantially 20 to 1 or greater.
engine. The split-cycle engine is disclosed generally in U.S.
In an alternative embodiment of the invention the expanPat. No. 6,543,225 to Scuderi, titled Split Four Stroke Internal 15
sionpiston and the compression piston of the engine have a
Combustion Engine, filed on Jul. 20, 2001, which is herein
TDC phasing of substantially 50° crank angle or less.
incorporated by reference in its entirety.
In another alternative embodiment of the invention, an
An exemplary embodiment of the split-cycle engine conengine includes a crankshaft, rotating about a crankshaft axis
cept is shown generally at 70. The split-cycle engine 70
replaces two adjacent cylinders of a conventional four-stroke 20 of the engine. An expansion piston is slidably received within
an expansion cy Iinder and operatively connected to the crankengine with a combination of one compression cylinder 72
shaft such that the expansion piston reciprocates through an
and one expansion cylinder 74. These two cylinders 72, 74
expansion stroke and an exhaust stroke of a four stroke cycle
would perform their respective functions once per crankshaft
during a single rotation of the crankshaft. A compression
76 revolution. The intake charge would be drawn into thecompression cylinder 72 through typical poppet-style valves 25 piston is slidably received within a compression cylinder and78. The compression cylinder piston 73 would pressurize theoperatively connected to the crankshaft such that the comcharge and drive the charge through the crossover passage 80,
pression piston reciprocates through an intake stroke and awhich acts as the intake port for the expansion cylinder 74. A
compression stroke of the same four stroke cycle during thecheck valve 82 at the inlet would be used to prevent reverse
same rotation of the crankshaft. A crossover passage interflow from the crossover passage 80. Valve(s) 84 at the outlet 30 connects the compression and expansion cylinders. Theof the crossover passage 80 would control the flow of thecrossover passage includes an inlet valve and a crossoverpressurized intake charge into the expansion cylinder 74.
valve defining a pressure chamber therebetween. The crossSpark plug 86 would be ignited soon after the intake chargeover valve has a crossover valve duration of substantially 69°
enters the expansion cylinder 74, and the resulting combusof crank angle or less.
tion would drive the expansion cylinder piston 75 down. 35
In still another embodiment of the invention an engine
Exhaust gases would be pumped out of the expansion cylinincludes a crankshaft, rotating about a crankshaft axis of the
der through poppet valves 88.
engine. An expansion piston is slidably received within an
With the split-cycle engine concept, the geometric engine
expansion cylinder and operatively connected to the crankparameters (i.e., bore, stroke, connecting rod length, Compression Ratio, etc.) of the compression and expansion cyl- 40 shaft such that the expansion piston reciprocates through an
expansion stroke and an exhaust stroke of a four stroke cycle
inders are generally independent from one another. For
during a single rotation of the crankshaft. A compression
example, the crank throws 90, 92 for each cylinder may havepiston is slidably received within a compression cylinder anddifferent radii and be phased apart from one another with top
operatively connected to the crankshaft such that the comdead center (TDC) of the expansion cylinder piston 75 occurring prior to TDC of the compression cylinder piston 73. This 45 pression piston reciprocates through an intake stroke and a
independence enables the split-cycle engine to potentiallycompression stroke of the same four stroke cycle during theachieve higher efficiency levels than the more typical foursame rotation of the crankshaft. A crossover passage interstroke engines previously described herein.
connects the compression and expansion cylinders. The
However, there are many geometric parameters and comcrossover passage includes an inlet valve and a crossoverbinations of parameters in the split-cycle engine. Therefore, 50 valve defining a pressure chamber therebetween. The crossfurther optimization of these parameters is necessary to maxiover valve remains open during at least a portion of a commize the performance of the engine.
bustion event in the expansion cylinder.
Accordingly, there is a need for an improved four strokeinternal combustion engine, which can enhance efficiency
BRIEF DESCRIPTION OF THE DRAWINGS
55
and reduce NOx emission levels.
FIG. 1 is a schematic diagram of a prior art conventional
four stroke internal combustion engine during the intake
stroke;The present invention offers advantages and alternatives
FIG. 2 is a schematic diagram of the prior art engine ofFI G.
over the prior art by providing a split-cycle engine in which 60
1duringthe compression stroke;significant parameters are optimized for greater efficiency
FIG. 3 is a schematic diagram of the prior art engine ofFI G.
and performance. The optimized parameters include at least
1 during the expansion stroke;
one of Expansion Ratio, Compression Ratio, top dead center
FIG. 4 is a schematic diagram of the prior art engine ofFI G.
phasing, crossover valve duration, and overlap between the
65 1 during the exhaust stroke;crossover valve event and combustion event.
FIG. 5 is a schematic diagram of a prior art split-cycle four
These and other advantages are accomplished in an exemstroke internal combustion engine;plary embodiment of the invention by providing an engine
SUMMARY OF THE INVENTION

US 8,006,656 B2

5

6FIG. 6 is a schematic diagram of an exemplary embodiFIG. 29 is a graph of indicated power and thermal effiment of a split-cycle four stroke internal combustion engine
ciency vs. crossover valve duration for a split cycle engine inaccordance with the present invention.
in accordance with the present invention during the intake
stroke;DETAILED DESCRIPTION
FIG. 7 is a schematic diagram of the split-cycle engine of
FIG. 6 during partial compression of the compression stroke;
I. Overview
FIG. 8 is a schematic diagram of the split-cycle engine of
FIG. 6 during full compression of the compression stroke;
The Scuderi Group, LLC commissioned the Southwest
FIG. 9 is a schematic diagram of the split-cycle engine of
10 Research Institute® (SwRI®) of San Antonio, Tex. to perFIG. 6 during the start of the combustion event;
form a Computerized Study. The Computerized Study
FIG. 9A is a schematic diagram similar to FIG. 9 illustratinvolved constructing a computerized model that represented
ing addition of fuel to the exit end of the crossover passage;
various embodiments of a split-cycle engine, which was comFIG. 10 is a schematic diagram of the split-cycle engine of
15 pared to a computerized model of a conventional four stroke
FIG. 6 during the expansion stroke;internal combustion engine having the same trapped mass per
FIG. 11 is a schematic diagram of the split-cycle engine ofcycle. The Study's final report (SwRI® Project No. 03.05932,
FIG. 6 during the exhaust stroke;
dated Jun. 24, 2003, titled "Evaluation Of Split-Cycle FourFIG. 12A is a schematic diagram of a GT-Power graphical
Stroke Engine Concept") is herein incorporated by referenceuser interface for a conventional engine computer model usedin
its entirety. The Computerized Study resulted in the present
in a comparative Computerized Study;
20
invention described herein through exemplary embodiments
FIG.12B is the item definitions of the conventional engine
pertaining to a split-cycle engine.
ofFIG.12A;
FIG. 13 is a typical Wiebe heat release curve;
II. Glossary
FIG. 14 is a graph of performance parameters ofthe con25
ventional engine of FIG. 12A;
The following glossary of acronyms and definitions of
FIG. 15A is a schematic diagram of a GT-Power graphicalterms
used herein is provided for reference:user interface for a split-cycle engine computer model in
Air/fuel Ratio: proportion of air to fuel in the intake charge
accordance with the present invention and used in the ComBottom Dead Center (BDC): the piston's farthest position
puterized Study;
30
from the cylinder head, resulting in the largest combustion
FIG. 15B is the item definitions of the split-cycle engine ofchamber volume of the cycle.
FIG.15A
Brake Mean Effective Pressure (BMEP): the engine's brake
FIG. 16 is a schematic representation of an MSC.ADtorque output expressed in terms of a MEP value. Equal to
AMS® model diagram of the split cycle engine of FIG. 15A;
the brake torque divided by engine displacement.
FIG. 17 is a graph of the compression and expansion piston 35 Brake Power: the power output at the engine output shaft.
positions and valve events for the split-cycle engine of FIG.
Brake Thermal Efficiency (BTE): the prefix "brake": having
15A;
to do with parameters derived from measured torque at the
FIG. 18 is a graph of some of the initial performanceengine output shaft. This is the performance parameterparameters of the split-cycle engine of FIG. 15A;
taken after the losses due to friction. Accordingly
FIG. 19 is a log-log pressure volume diagram for a con- 40
BTE=ITE-friction.
ventional engine;
Bum Overlap: the percentage of the total combustion event
FIG. 20 is a pressure volume diagram for the power cylin(i.e. from the 0% point to the 100% point of combustion)that is completed by the time of crossover valve closing.
der of a split-cycle engine in accordance with the present
invention;
Brake Torque: the torque output at the engine output shaft.
FIG. 21 is a comparison graph of indicated thermal effi- 45 Crank Angle (CA): the angle of rotation of the crankshaft
ciencies of a conventional engine and various split-cycle
throw, typically referred to its position when aligned withengines in accordance with the present invention;
the cylinder bore.
Computational Fluid Dynamics (CFD): a way of solving
FIG. 22 is a CFD predicted diagram of the flame front
position between the crossover valve and expansion piston forcomplex fluid flow problems by breaking the flow regime
up into a large number of tiny elements which can then be50
a 35% bum overlap case;solved to determine the flow characteristics, the heat transFIG. 23 is a CFD predicted diagram of the flame front
fer and other characteristics relating to the flow solution.
position between the crossover valve and expansion piston for
Carbon Monoxide (CO): regulated pollutant, toxic to
a 5% bum overlap case;humans, a product of incomplete oxidation ofhydrocarbon
FIG. 24 is a CFD predicted graph ofNOx emissions for aconventional engine, a split-cycle engine 5% burn overlap 55
fuels.
case and a split-cycle engine 35% bum overlap case;
Combustion Duration: defined for this text as the crank angle
interval between the 10% and 90% points from the start of
FIG. 25 is a graph of the expansion piston thrust load for thethe combustion event. Also known as the Bum Rate. See
split-cycle engine;
the Wiebe Heat Release Curve in FIG. 13.
FIG. 26 is a graph of indicated power and thermal efficiency vs. Compression Ratio for a split-cycle engine in 60 Combustion Event: the process of combusting fuel, typicallyaccordance with the present invention;in the expansion chamber of an engine.
FIG. 27 is a graph of indicated power and thermal effiCompression Ratio: ratio of compression cylinder volume atciency vs. Expansion Ratio for a split cycle engine in accorBDC to that at TDC
dance with the present invention;
Crossover Valve Closing (XVC)
FIG. 28 is a graph of indicated power and thermal effi- 65 Crossover Valve Opening (XVO)
Cylinder Offset: is the linear distance between a bore's cenciency vs. TDC phasing for a split cycle engine in accordance
terline and the crankshaft axis.
with the present invention; andUS 8,006,656 B2
7
Displacement Volume: is defined as the volume that the piston
displaces from BDC to TDC. Mathematically, if the stroke
is defined as the distance from BDC to TDC, then thedisplacement volume is equal to n/4*bore 2 *stroke. Compression Ratio is then the ratio of the combustion chamber
volume at BDC to that at TDC. The volume at TDC isreferred to as the clearance volume, or V c1·

8

From these definitions, it follows that:Net Indicated=Gross Indicated+ Pumping.
Brake=Net Indicated-Friction
Pumping Mean Effective Pressure (PMEP): the indicated
MEP associated with just the exhaust and intake strokes. A
5measure of power consumed in the breathing process.
However, sign convention taken is that a positive valuemeans that work is being done on the crankshaft during thepumping loop. (It is possible to get a positive value for
Vf"Jt/4*bore2 *stroke
10
PMEP if the engine is turbocharged or otherwise boosted.)
Spark-Ignited (SI): refers to an engine in which the combustion event is initiated by an electrical spark inside the
combustion chamber.
Exhaust Valve Closing (EVC)
Top Dead Center (TDC): the closest position to the cylinder
Exhaust Valve Opening (EVO)
Expansion Ratio: is the equivalent term to Compression 15 head that the piston reaches throughout the cycle, providing
the lowest combustion chamber volume.
Ratio, but for the expansion cylinder. It is the ratio of
TDC Phasing (also referred to herein as the phase angle
cylinder volume at BDC to the cylinder volume at TDC.
between the compression and expansion cylinders (seeFriction Mean Effective Pressure (FMEP): friction level
item 172 of FIG. 6)): is the rotational offset, in degrees,
expressed in terms of a MEP. Cannot be determined
20
between the crank throw for the two cylinders. A zero
directly from a cylinder pressure curve though. One comdegree offset would mean that the crank throws were comon way of measuring this is to calculate the NIMEP fromlinear, while a 180° offset would mean that they were on
the cylinder pressure curve, calculate the BMEP from the
opposite sides of the crankshaft (i.e. one pin at the top while
torque measured at the dynamometer, and then assign thethe other is at the bottom).
difference as friction or FMEP.
25 Thermal Efficiency: ratio of power output to fuel energy input
Graphical User Interface (GUI)rate. This value can be specified as brake (BTE) or indiIndicated Mean Effective Pressure (IMEP): the integration ofcated (ITE) thermal efficiency depending on which power
the area inside the P-dV curve, which also equals the indiparameter is used in the numerator.
cated engine torque divided by displacement volume. In
VP: mean piston velocity: the average velocity of the piston
fact, all indicated torque and power values are derivatives 30
throughout the cycle. Can be expressed mathematically asof this parameter. This value also represents the constant
2*Stroke*Engine Speed.
pressure level through the expansion stroke that would
Valve Duration (or Valve Event Duration): the crank angle
provide the same engine output as the actual pressureinterval between a valve opening and a valve closing.
curve. Can be specified as net indicated (NIMEP) or gross
Valve Event: the process of opening and closing a valve toindicated (GIMEP) although when not fully specified, 35
perform a task.
NIMEP is assumed.
Volumetric Efficiency: the mass of charge (air and fuel)Indicated Thermal Efficiency (ITE): the thermal efficiency
trapped in the cylinder after the intake valve is closed
based on the (net) indicated power.
compared to the mass of charge that would fill the cylinder
Intake Valve Closing (IVC)
displacement volume at some reference conditions. The
40
Intake Valve Opening (IVO)
reference conditions are normally either ambient, or intake
Mean Effective Pressure: the pressure that would have to be
manifold conditions. (The latter is typically used on turboapplied to the piston through the expansion stroke to result
charged engines.)in the same power output as the actual cycle. This value is
Wide-Open Throttle (WOT): refers to the maximum achievalso proportional to torque output per displacement.
able output for a throttled (SI) engine at a given speed.
NOx: various nitrogen oxide chemical species, chiefly NO 45
and N0 2 . A regulated pollutant and a pre-cursor to smog.
III. Embodiments of the Split-Cycle Engine
Created by exposing an environment including oxygen and
Resulting from the Computerized Study
nitrogen (i.e. air) to very high temperatures.
Referring to FIGS. 6-11, an exemplary embodiment of a
Peak Cylinder Pressure (PCP): the maximum pressure
achieved inside the combustion chamber during the engine 50 four stroke internal combustion engine in accordance with thecycle.
present invention is shown generally at 100. The engine 100
Prefixes:-Power, Torque, MEP, Thermal Efficiency and otherincludes an engine block 102 having an expansion (or power)cy Iinder 104 and a compression cy Iinder 106 extending thereterms may have the following qualifying prefixes:
Indicated: refers to the output as delivered to the top of thethrough. A crankshaft 108 is pivotally connected for rotation
55 about a crankshaft axis 110 (extending perpendicular to thepiston, before friction losses are accounted for.
plane of the paper).
Gross Indicated: refers to the output delivered to the top of
the piston, considering only compression and expansion
The engine block 102 is the main structural member of thestrokes.
engine 100 and extends upward from the crankshaft 108 to the
Net Indicated: (also the interpretation of"indicated" whenjunction with a cylinder head 112. The engine block 102
not otherwise denoted): refers to the output delivered to 60 serves as the structural framework of the engine 100 andthe top of the piston considering all four strokes of thetypically carries the mounting pad by which the engine iscycle: compression, expansion, exhaust, and intake.
supported in the chassis (not shown). The engine block 102 is
generally a casting with appropriate machined surfaces andPumping: refers to the output of the engine considering
threaded holes for attaching the cylinder head 112 and other
only the intake and exhaust strokes. In this report, positive pumping work refers to work output by the engine 65 units of the engine 100.
while negative relates to work consumed by the engine
The cylinders 104 and 106 are openings of generally cirto perform the exhaust and intake strokes.
cular cross section, that extend through the upper portion ofUS 8,006,656 B2
910

the engine block 102. The diameter of the cylinders 104 andover the top to the first cylinder 104. Poppet valves 150, 152
106 is known as the bore. The internal walls of cylinders 104
and 154 typically have a metal shaft (or stem) 156 with a disk
and 106 are bored and polished to form smooth, accurate
158 at one end fitted to block the valve opening. The other end
bearing surfaces sized to receive an expansion (or power)of the shafts 156 of poppet valves 150, 152 and 154 are
piston 114, and a compression piston 116 respectively.
5 mechanically linked to camshafts 160, 162 and 164 respecThe expansion piston 114 reciprocates along an expansion
tively. The camshafts 160, 162 and 164 are typically a round
piston-cylinder axis 113, and the compression piston 116
rod with generally oval shaped lobes located inside the engine
reciprocates along a second compression piston-cylinder axis
block 102 or in the cylinder head 112.
115. In this embodiment, the expansion and compression
The camshafts 160, 162 and 164 are mechanically concylinders 104 and 106 are offset relative to crankshaft axis
10 nected to the crankshaft 108, typically through a gear wheel,
110. That is, the first and second piston-cylinder axes 113 andbelt or chain links (not shown). When the crankshaft 108
115 pass on opposing sides of the crankshaft axis 110 without
forces the camshafts 160, 162 and 164 to turn, the lobes on theintersecting the crankshaft axis 110. However, one skilled incamshafts 160,162 and 164 cause the valves 150,152 and154
the art will recognize that split-cycle engines without offset
piston-cylinder axis are also within the scope of this inven- 15 to open and close at precise moments in the engine's cycle.
The crown 120 of compression piston 116, the walls oftion.
second cylinder 106 and the cylinder head 112 form a comThe pistons 114 and 116 are typically cylindrical castingspression chamber 166 for the second cy Iinder 106. The crownor forgings of steel or aluminum alloy. The upper closed ends,
118 of power piston 114, the walls of first cylinder 104 and thei.e., tops, of the power and compression pistons 114 and 116
are the first and second crowns 118 and 120 respectively. The 20 cylinder head 112 form a separate combustion chamber 168
for the first cylinder 104. A spark plug 170 is disposed in theouter surfaces of the pistons 114, 116 are generally machined
cylinder head 112 over the first cylinder 104 and is controlled
to fit the cylinder bore closely and are typically grooved toby a control device (not shown) which precisely times thereceive piston rings (not shown) that seal the gap between theignition of the compressed air gas mixture in the combustion
pistons and the cylinder walls.
First and second connecting rods 122 and 124 are pivotally 25 chamber 168.
attached at their top ends 126 and 128 to the power andThough this embodiment describes a spark ignition (SI)compression pistons 114 and 116 respectively. The crankengine, one skilled in the art would recognize that compresshaft 108 includes a pair of mechanically offset portions
sion ignition (CI) engines are within the scope of this type ofengine also. Additionally, one skilled in the art would recogcalled the first and second throws 130 and 132, which arepivotally attached to the bottom opposing ends 134 and 136 of 30 nize that a split-cycle engine in accordance with the present
the first and second connecting rods 122 and 124 respectively.
invention can be utilized to run on a variety of fuels other thanThe mechanical linkages of the connecting rods 122 and 124
gasoline, e.g., diesel, hydrogen and natural gas.
to the pistons 114, 116 and crankshaft throws 130, 132 serve
During operation the power piston 114 leads the compresto convert the reciprocating motion of the pistons (as indision piston 116 by a phase angle 172, defined by the degrees
cated by directional arrow 138 for the expansion piston 114, 35 of crank angle (CA) rotation the crankshaft 108 must rotate
after the power piston 114 has reached its top dead centerand directional arrow 140 for the compression piston 116) to
the rotary motion (as indicated by directional arrow 142) ofposition in order for the compression piston 116 to reach its
respective top dead center position. As will be discussed in thethe crankshaft 108.
Computer Study hereinafter, in order to maintain advantaThough this embodiment shows the first and second pistons 114 and 116 connected directly to crankshaft 108 40 geous thermal efficiency levels (BTE or ITE), the phase angle
through connecting rods 122 and 124 respectively, it is within
172 is typically set at approximately 20 degrees. Moreover,the scope of this invention that other means may also be
the phase angle is preferably less than or equal to 50 degrees,more preferably less than or equal to 30 degrees and mostemployed to operatively connect the pistons 114 and 116 topreferably less than or equal to 25 degrees.
the crankshaft 108. For example a second crankshaft may be
used to mechanically link the pistons 114 and 116 to the first 45
FIGS. 6-11 represent one full cycle of the split cycle engine
crankshaft 108.
100 as the engine 100 converts the potential energy of a
The cylinder head 112 includes a gas crossover passage
predetermined trapped mass of air/fuel mixture (represented
144 interconnecting the first and second cylinders 104 andby the dotted section) to rotational mechanical energy. That
106. The crossover passage includes an inlet check valve 146
is, FIGS. 6-11 illustrate intake, partial compression, full comdisposed in an end portion of the crossover passage 144 so pression, start of combustion, expansion and exhaust of theproximate the second cylinder 106. A poppet type, outlet
trapped mass respectively. However, it is important to notecrossover valve 150 is also disposed in an opposing end
that engine is fully charged with air/fuel mixture throughout,portion of the crossover passage 144 proximate the top of theand that for each trapped mass of air/fuel mixture taken in andfirst cylinder 104. The check valve 146 and crossover valve
compressed through the compression cylinder 106, a substan150 define a pressure chamber 148 there between. The check 55 tially equal trapped mass is combusted and exhausted throughvalve 146 permits the one way flow of compressed gas fromthe expansion cylinder 104.
FIG. 6 illustrates the power piston 114 when it has reached
the second cylinder 106 to the pressure chamber 148. The
crossover valve 150 permits the flow of compressed gas fromits bottom dead center (BDC) position and has just started
the pressure chamber 148 to the first cylinder 104. Thoughascending (as indicated by arrow 138) into its exhaust stroke.
check and poppet type valves are described as the inlet check 60 Compression piston 116 is lagging the power piston 114 andis descending (arrow 140) through its intake stroke. The inlet
and the outlet crossover valves 146 and 150 respectively, any
valve design appropriate for the application may be usedvalve 152 is open to allow a predetermined volume of explosive mixture of fuel and air to be drawn into the compression
instead, e.g., the inlet valve 146 may also be of the poppet
type.
chamber 166 and be trapped therein (i.e., the trapped mass as
The cylinder head 112 also includes an intake valve 152 of 65 indicated by the dots on FIG. 6). The exhaust valve 154 is also
the poppet type disposed over the top of the second cylinder
open allowing piston 114 to force spent products of combus106, and an exhaust valve 154 of the poppet type disposedtion out of the combustion chamber 168.

US 8,006,656 B2
11

12

The check valve 146 and crossover valve 150 of the crosssmall compared to the valve duration of the intake valve 152
overpassage 144 are closed to prevent the transfer of ignitableand exhaust valve 154. A typical valve duration for valves 152
fuel and spent combustion products between the two chamand 154 is typically in excess of 160 degrees CA. In order tobers 166 and 168. Additionally during the exhaust and intake
maintain advantageous efficiency levels, the crossover valve
strokes, the check valve 146 and crossover valve 150 seal the
duration is typically set at approximately 25 degrees CA.
pressure chamber 148 to substantially maintain the pressure
Moreover, the crossover valve duration is preferably equal toof any gas trapped therein from the previous compression andor less than 69 degrees CA, more preferably equal to or lesspower strokes.
than 50 degrees CA, and most preferably equal to or less thanReferring to FIG. 7, partial compression of the trapped
35 degrees CA.
mass is in progress. That is inlet valve 152 is closed and 10
Additionally, the Computer Study also indicated that if thecompression piston 116 is ascending (arrow 140) toward its
crossover valve duration and the combustion duration overtop dead center (TDC) position to compress the air/fuel mixlapped by a predetermined minimum percentage of combusture. Simultaneously, exhaust valve 154 is open and thetion duration, then the combustion duration would be subexpansion piston 114 is also ascending (arrow 138) to exhauststantially decreased (that is the bum rate of the trapped massspent fuel products.
15 would be substantially increased). Specifically, the crossover
Referring to FIG. 8, the trapped mass (dots) is further
valve 150 should remain open preferably for at least 5% ofcompressed and is beginning to enter the crossover passage
the total combustion event (i.e. from the 0% point to the 100%
144 through check valve 146. The expansion piston 114 haspoint of combustion) prior to crossover valve closing, morereached its top dead center (TDC) position and is about topreferably for 10% of the total combustion event, and mostdescend into its expansion stroke (indicated by arrow 138), 20 preferably for 15% of the total combustion event. As
while the compression piston 116 is still ascending throughexplained in greater detail hereinafter, the longer the crossits compression stroke (indicated by arrow 140).At this point,
over valve 150 can remain open during the time the air/fuelcheck valve 146 is partially open. The crossover outlet valve
mixture is combusting (i.e., the combustion event), the greater
150, intake valve 152 and exhaust valve 154 are all closed.
the increase in bum rate and efficiency levels will be. LimiAt TDC piston 114 has a clearance distance 178 between 25 tations to this overlap will be discussed in later sections.
the crown 118 of the piston 114 and the top of the cylinder
Upon further rotation of the crankshaft 108, the compres104. This clearance distance 178 is very small by comparison
sion piston 116 will pass through to its TDC position andto the clearance distance 60 of a conventional engine 10 (bestthereafter start another intake stroke to begin the cycle overseen in prior art FIG. 3). This is because the clearance (oragain. The compression piston 116 also has a very small
Compression Ratio) on the conventional engine is limited to 30 clearance distance 182 relative to the standard engine 10. Thisavoid inadvertent compression ignition and excessive cylinis possible because, as the gas pressure in the compression
der pressure. Moreover, by reducing the clearance distancechamber 166 of the compression cylinder 106 reaches the178, a more thorough flushing of the exhaust products ispressure in the pressure chamber 148, the check valve 146 isaccomplished.
forced open to allow gas to flow through. Therefore, a very
The ratio of the expansion cylinder volume (i.e., combus- 35 small volume of high pressure gas is trapped at the top of thetion chamber 168) when the piston 114 is at BDC to thecompression piston 116 when it reaches its TDC position.
expansion cylinder volume when the piston is at TDC is
The ratio of the compression cylinder volume (i.e., comdefined herein as the Expansion Ratio. This ratio is generallypression chamber 166) when the piston 116 is at BDC to themuch higher than the ratio of cylinder volumes between BDC
compression cylinder volume when the piston is at TDC is
and TDC of the conventional engine 10. As indicated in the 40 defined herein as the Compression Ratio. This ratio is generfollowing Computer Study description, in order to maintainally much higher than the ratio of cylinder volumes between
advantageous efficiency levels, the Expansion Ratio is typiBDC and TDC of the conventional engine 10. As indicated incally set at approximately 120 to 1. Moreover, the Expansion
the following Computer Study description, in order to mainRatio is preferably equal to or greater than 20 to 1, moretain advantageous efficiency levels, the Compression Ratio ispreferably equal to or greater than 40 to 1, and most prefer- 45 typically set at approximately 100 to 1. Moreover, the Comably equal to or greater than 80 to 1.
pression Ratio is preferably equal to or greater than 20 to 1,
Referring to FIG. 9, the start of combustion of the trapped
more preferably equal to or greater than 40 to 1, and mostmass (dotted section) is illustrated. Addition of fuel to the exit
preferably equal to or greater than 80 to 1.
end of the crossover passage 144 by a fuel injection system
Referring to FIG. 10, the expansion stroke on the trapped
149 is illustrated in FIG. 9A. The crankshaft 108 has rotated 50 mass is illustrated. As the air/fuel mixture is combusted, thean additional predetermined number of degrees past the TDC
hot gases drive the expansion piston 114 down.
position of expansion piston 114 to reach its firing position.
Referring to FIG. 11, the exhaust stroke on the trapped
At this point, spark plug 170 is ignited and combustion ismass is illustrated. As the expansion cylinder reaches BDC
started. The compression piston 116 is just completing its
and begins to ascend again, the combustion gases arecompression stroke and is close to its TDC position. During 55 exhausted out the open valve 154 to begin another cycle.
this rotation, the compressed gas within the compression
IV. Computerized Studycylinder 116 reaches a threshold pressure which forces thecheck valve 146 to fully open, while cam 162 is timed to also
open crossover valve 150. Therefore, as the power piston 114
1.0 Summary of Results
descends and the compression piston 116 ascends, a sub stan- 60
tially equal mass of compressed gas is transferred from the
1.1. Advantages
compression chamber 166 of the compression cylinder 106 to
The primary objective of the Computerized Study was tothe combustion chamber 168 of the expansion cylinder 104.
study the concept split-cycle engine, identifY the parameters
exerting the most significant influence on performance andAs noted in the following Computer Study description, it isadvantageous that the valve duration of crossover valve 150, 65 efficiency, and determine the theoretical benefits, advantages,i.e., the crank angle interval (CA) between the crossover valve
or disadvantages compared to a conventional four-stroke
opening (XVO) and crossover valve closing (XVC), be very
engine.

US 8,006,656 B2
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14

The Computerized Study identified Compression Ratio,as the efficiency of the engine. It is a well known fact that
Expansion Ratio, TDC phasing (i.e., the phase angle between
efficiencies can be improved on SI engines by running lean(significantly above 14.5 to 1 air/fuel ratio). However, thethe compression and expansion pistons (see item 172 of FIG.
dependence on three way catalytic converters (TWC), which6)), crossover valve duration and combustion duration as
require a stoichiometric exhaust stream in order to reachsignificant variables affecting engine performance and effirequired emissions levels, typically precludes this option on
ciency. Specifically the parameters were set as follows:
production engines. (Stoichiometric air/fuel ratio is about
the compression and Expansion Ratios should be equal to14.5 for gasoline fuel.) The lower NOx emissions of the splitor greater than 20 to 1 and were set at 100 to 1 and 120
to 1 respectively for this Study;
cycle engine may allow the split-cycle to run lean and achieve
the phase angle should be less than or equal to 50 degrees 10 additional efficiency gains on the order of one point (i.e.,
and was set at approximately 20 degrees for this study;
approximately 3%) over a conventional engine with a conand
ventional TWC. TWCs on conventional engines demonstrate
NOx reduction levels of above 95%, so the split-cycle engine
the crossover valve duration should be less than or equal to
69 degrees and was set at approximately 25 degrees for
carmot reach their current post-TWC levels, but depending on
15 the application and with the use of other aftertreatment techthis Study.
nology, the split-cycle engine may be able to meet required
Moreover, the crossover valve duration and the combustion
duration should overlap by a predetermined percent of theNOx levels while running at lean air/fuel ratios.
combustion event for enhanced efficiency levels. For this
These results have not been correlated to experimental
Study, CFD calculations showed that an overlap of 5% of thedata, and emissions predictions from numerical models tend
total combustion event was realistic and that greater overlaps 20 to be highly dependent on tracking of trace species throughare achievable with 35% forming the unachievable upper
the combustion event. If these results were confirmed on anlimit for the embodiments modeled in this study.
actual test engine, they would constitute a significant advanWhen the parameters are applied in the proper configuratage of the split-cycle engine concept.
tion the split-cycle engine displayed significant advantages in
1.2 Risks And Suggested Solutions:both brake thermal efficiency (BTE) and NOx emissions. 25
The Computerized Study also identified the following risks
Table 9 summarized the results of the Computerized Study
associated with the split-cycle engine:with regards to BTE, and FIG. 24 graphs the predicted NOx
Sustained elevated temperatures in the expansion cylinder
emissions, for both the conventional engine model and varicould lead to thermal-structural failures of components
ous embodiments of the split-cycle engine model.
and problems with lube oil retention,
The predicted potential gains for the split-cycle engine 30
Possible valve train durability issues with crossover valve
concept at the 1400 rpm engine speed are in the range of0.7
due to high acceleration loads,
to less than 5.0 points (or percentage points) of brake thermal
Valve-to-piston interference in the expansion cylinder, andefficiency (BTE) as compared to that of a conventional four
Auto-ignition and/or flame propagation into crossover passtroke engine at 33.2 points BTE. In other words, the BTE ofsage.
However, the above listed risks may be addressed throughthe split-cycle engine was calculated to be potentially 35
a myriad of possible solutions. Examples of potential techbetween 33.9 and 38.2 points.
The term "point" as used herein, refers to the absolute
nologies or solutions that may be utilized are given below.
Dealing with the sustained high temperatures in the expancalculated or measured value of percent BTE out of a theoretically possible 100 percentage points. The term "percent",
sian cylinder may utilize unique materials and/or construcas used herein, refers to the relative comparative difference 40 tion techniques for the cylinder wall. In addition, lower tembetween the calculated BTE of the split-cycle engine and theperature and/or different coolants may need to be used. Alsoof concern in dealing with the high temperatures is the lubribase line conventional engine. Accordingly, the range of0.7
cation issue. Possible technologies for overcoming this chalto less than 5.0 points increase in BTE for the split-cycle
lenge are extreme high temperature-capable liquid lubricants
engine represents a range of approximately 2 (i.e., 0.7/33.2)
to less than 15 (5/33.2) percent increase in BTE over the 45 (advanced synthetics) as well as solid lubricants.
baseline of 33.2 for a conventional four stroke engine.
Addressing the second item of valvetrain loads for the very
Additionally, the Computerized Study also showed that if
quick-acting crossover valve may include some of the techthe split-cycle engine were constructed with ceramic expannology currently being used in advanced high speed racing
sion piston and cylinder, the BTE may potentially further
engines such as pneumatic valve springs and/or low inertia,
increase by as much as 2 more points, i.e., 40.2 percentage 50 titanium valves with multiple mechanical springs per valve.
points BTE, which represents an approximate 21 percent
Also, as the design moves forward into detailed design, theincrease over the conventional engine. One must keep in mind
number of valves will be reconsidered, as it is easier to movea larger number of smaller valves more quickly and they
however, that ceramic pistons and cylinders have durability
problems with long term use; in addition, this approach would
provide a larger total circumference providing better flow atfurther aggravate the lubrication issues with the even higher 55 low lift.
temperature cylinder walls that would result from the use of
The third item of crossover valve interference with thethese materials.
piston near TDC may be addressed by recessing the crossoverWith the stringent requirements on emissions and the marvalves in the head, providing reliefs or valve cutouts in the
ket need for increased efficiency, many engine manufacturers
piston top to allow space for the valve(s ), or by designing an
struggle to reduce NOx emissions while operating at lean 60 outward-opening crossover valve.
The last challenge listed is auto-ignition and/or flame
air/fuel ratios. An output of a CFD combustion analysis perpropagation into the crossover passage. Auto-ignition in theformed during the Computer Study indicated that the splitcrossover passage refers to the self-ignition of the air/fuelcycle engine could potentially reduce the NOx emissions levels of the conventional engine by 50% to 80% when
mixture as it resides in the crossover passage between cycles
65 due to the presence of a combustible mixture held for acomparing both engines at a lean air/fuel ratio.
relatively long duration at high temperature and pressure.
The reduction in NOx emissions could potentially be significant both in terms of its impact on the environment as wellThis can be addressed by using port fuel injection, where onlyUS 8,006,656 B2

15

16

air resides in the crossover passage between cycles therefore
2.2 Conventional Engine Model Construction
preventing auto-ignition. The fuel is then added either
The engine characteristics were selected to be representadirectly into the cylinder, or to the exit end of the crossover
tive of typical gasoline SI engines. The engine displacement
passage, timed to correspond with the crossover valve openwas similar to a two-cylinder version of an automotive application in-line four-cylinder 202 in3 (3.3 L) engine. The Comingtime.
pression Ratio was set to 8.0:1. The stoichiometric air/fuelThe second half of this issue, flame propagation into theratio for gasoline, which defines the proportions of air andcrossover passage, can be further optimized with development. That is, although it is very reasonable to design thefuel required to convert all of the fuel into completely oxitiming of the split-cycle engine's crossover valve to be open
dized products with no excess air, is approximately 14.5:1.
10
The selected air/fuel ratio of 18: 1 results in lean operation.
during a small portion of the combustion event, e.g., 5% orTypical automotive gasoline SI engines operate at stoichioless, the longer the crossover valve is open during the commetric or slightly rich conditions at full load. However, leanbustion event the greater the positive impact on thermal effioperation typically results in increased thermal efficiency.
ciency that can be achieved in this engine. However, thisThe typical gasoline SI engine runs at stoichiometric condirection of increased overlap between the crossover valve 15
ditions because that is a requirement for proper operation ofand combustion events increases the likelihood of flame
the three-way catalytic converter. The three-way catalyst
propagation into the cross-over passage. Accordingly, effort(TWC) is so-named due to its ability to provide both the
can be directed towards understanding the relationshipoxidation of HC and CO to H 2 0 and C0 2 , as well as the
between combustion timing, spark plug location, crossover
reduction ofNOx to N 2 and 0 2 . These TWCs are extremely
valve overlap and piston motion in regards to the avoidance of 20
effective, achieving reductions of over 90% of the incoming
flame propagation into the crossover passage.
pollutant stream but require close adherence to stoichiometric
operation. It is a well known fact that efficiencies can be
2.0 Conventional Engine Model
improved on SI engines by running lean, but the dependence
on TWCs to reach required emissions levels typically preA cycle simulation model was constructed of a two-cylin- 25 eludes this option on production engines.
der conventional naturally-aspirated four-stroke SI engine
It should be noted that under lean operation, oxidation
and analyzed using a commercially available software packcatalysts are readily available which will oxidize HC and CO,
age called GT-Power, owned by Gamma Technologies, Inc.
but reduction ofNOx is a major challenge under such condiofWestmont, Ill. The characteristics of this model were tuned
tions. Developments in the diesel engine realm have recently
using representative engine parameters to yield performance 30 included the introduction of lean NOx traps and lean NOx
and efficiency values typical of naturally-aspirated gasoline
catalysts. At this point, these have other drawbacks such as
SI engines. The results from these modeling efforts were usedpoor reduction efficiency and/or the need for periodic regento establish a baseline of comparison for the split-cycle
eration, but are currently the focus of a large amount ofengine concept.
development.
2.1 GT-Power Overview
35
In any case, the major focus of the Computerized Study is
GT-Power is a 1-d computational fluids-solver that is comthe relative efficiency and performance. Comparing bothmonly used in industry for conducting engine simulations.
engines (split-cycle and conventional) at 18:1 air/fuel ratio
GT-Power is specifically designed for steady state and tranprovides comparable results. Either engine could be operated
sient engine simulations. It is applicable to all types of interinstead under stoichiometric conditions such that a TWC
nal combustion engines, and it provides the user with several 40 would function and both would likely incur similar performenu-based objects to model the many different components
mance penalties, such that the relative results of this study
that can be used on internal combustion engines. FIG. 12A
would still stand. The conventional engine parameters areshows the GT-Power graphical user interface (GUI) for thelisted in Table 1.
two-cylinder conventional engine model.
Referring to FIGS. 12A and B, Intake air flows from the 45
TABLE 1
ambient source into the intake manifold, represented by junctions 211 and 212. From there, the intake air enters the intake
Conventional Engine Parameters
ports (214-217) where fuel is injected and mixed with the
Value
Parameter
airstream. At the appropriate time of the cycle, the intake
valves (vix-y) open while the pistons in their respective cyl- 50
4.0 in (101.6 mm)
Bore
Stroke4.0 in (101.6 mm)
inders (cyl1 and cyl2) are on their downstroke (intake stroke).
Connecting Rod Length
9.6 in (243.8 mm)
The air and fuel mixture are admitted into the cylinder during
2.0 in (50.8 mm)
Crank Throw
this stroke, after which time the intake valves close. (Cyl 1
Displacement Volume
50.265 in3 (0.824 L)
and cyl2 are not necessarily in phase; i.e. they may go through
Clearance Volume
7.180 in 3 (0.118 L)
Compression Ratio
8.0:1
the intake process at completely different times.) After the 55
Engine Speed
1400 rpm
intake stroke, the piston rises and compresses the mixture to
Air/Fuel Ratio
18:1
a high temperature and pressure. Near the end of the compression stroke, the spark plug is energized which begins theInitially, the engine speed was set at 1400 rpm. This speed
burning of the air/fuel mixture. It bums, further raising thetemperature and pressure of the mixture and pushing down on 60 was to be used throughout the project for the parametric
the piston through the expansion or power stroke. Near the
sweeps. However, at various stages of the model construction,end of the expansion stroke, the exhaust valve opens and the
speed sweeps were conducted at 1400, 1800, 2400, and 3000
piston begins to rise, pushing the exhaust out of the cylinderrpm.
The clearance between the top of the piston and the cylininto the exhaust ports (229-232). From the exhaust ports, theexhaust is transmitted into the exhaust manifold (233-234) 65 der head was initially recommended to be 0.040 in (1 mm). To
and from there to the end environment (exhaust) representing
meet this requirement with the 7.180 in3 (0.118 L) clearance
the ambient.
volume would require a bawl-in-piston combustion chamber,US 8,006,656 B2

17

18

which is uncommon for automotive SI engines. More often,
start and end are difficult to ascertain due to their asymptotic
automotive SI engines feature pent-roof combustion chamapproach to the 0 and 100% burn lines. This is especially truebers. SwRI® assumed a flat-top piston and cylinder head towith respect to test data, where the heat release curve is a
simplifY the GT-Power model, resulting in a clearance of
calculated profile based on the measured cylinder pressure
0.571 in (14.3 mm) to meet the clearance volume requirecurve and other parameters. Therefore, the 10 and 90% bum
ment. There was a penalty in brake thermal efficiency (BTE)
points are used to represent the nominal "ends" of the heatof 0.6 points with the larger piston-to-head clearance.
release curve. In the Wiebe correlation, the user specifies the
The model assumes a four-valve cylinder head with two
duration of the 10-90% burn period (i.e. 10-90% duration)
1.260 in (32 mm) diameter intake valves and two 1.102 in (28and that controls the resultant rate of heat release. The usermm) diameter exhaust valves. The intake and exhaust ports 10
can also specifY the crank angle location of some other point
were modeled as straight sections of pipe with all flow losseson the profile, most typically either the 10 or 50% point, as anaccounted for at the valve. Flow coefficients at maximum list
anchor to provide the phasing of the heat release curve relawere approximately 0.57 for both the intake and exhaust,
tive to the engine cycle.
which were taken from actual flow test results from a repreThe wall temperature solver in GT-Power was used tosentative engine cylinder head. Flow coefficients are used to 15
predict
the piston, cylinder head, and cylinder liner wall temquantifY the flow performance of intake and exhaust ports onperatures for the conventional engine. GT-Power is continuengines. A 1.0 value would indicate a perfect port with noously calculating the heat transfer rates from the working
flow losses. Typical maximum lift values for real engine ports
fluid to the walls of each passage or component (including
are in the 0.5 to 0.6 range.
Intake and exhaust manifolds were created as 2.0 in (50.8 20 cylinders). This calculation needs to have the wall temperature as a boundary condition. This can either be provided as amm) diameter pipes with no flow losses. There was no throttle
fixed input, or the wall temperature solver can be turned on tomodeled in the induction system since the focus is on widecalculate it from other inputs. In the latter case, wall thickness
open throttle (WOT), or full load, operation. The fuel is delivand material are specified so that wall conductivity can beered via multi-port fuel injection.
The valve events were taken from an existing engine and 25 determined. In addition, the bulk fluid temperature that thescaled to yield realistic performance across the speed range
backside of the wall is exposed to is provided, along with the(1400, 1800, 2400 and 3000 rpm), specifically volumetricconvective heat transfer coefficient. From these inputs, theefficiency. Table 2 lists the valve events for the conventional
program solves for the wall temperature profile which is aengine.
function of the temperature and velocity of the working fluid,
30 among other things. The approach used in this work was thatthe wall temperature solver was turned on to solve for realistic
TABLE2temperatures for the cylinder components and then those
Conventional Engine Breathing and Combustion Parameterstemperatures were assigned to those components as fixed
temperatures for the remaining runs.
Parameter
Value
Cylinder head coolant was applied at 200° F. (366 K) with
35
28° BTDC-breathing
332° ATDC-firing
Intake Valve Openinga heat transfer coefficient of3000 W/m2 -K. The underside of(NO)
the piston is splash-cooled with oil applied at 250° F. (394 K)
557° ATDC-firing
Intake Valve Closing (IVC) l7°ABDC
with a heat transfer coefficient of 5 W/m 2 -K. The cylinder
Peak Intake Valve Lift
0.412 in (10.47 mm)
53° BBDC
127° ATDC-firing
Exhaust Valve Opening
walls are cooled via coolant applied at 200° F. (366 K) with a2(EVO)
40 heat transfer coefficient of 500 W/m -K and oil applied at
3 7° ATDC-breathing
397° ATDC-firing
Exhaust Valve Closing
F.
(394
K)
with
a
heat
transfercoefficientofl 000 W /m2 250°
(EVC)
K.
These
thermalboundary
conditions
wereapplied to thePeak Exhaust Valve Lift
0.362 in (9.18 mm)
10° ATDC-firing
10° ATDC-firing
50% Bum Point
model to predict the in-cylinder component surface tempera24° crank angle (CA)Combustion Durationtures. The predicted temperatures were averaged across the(10-90%)45 speed range and applied as fixed wall temperatures in theremaining simulations. Fixed surface temperatures for the
The combustion process was modeled using an empirical
piston 464° F. (513 K), cylinder head 448° F. (504 K), andWiebe heat release, where the 50% bum point and 10 to 90%
liner 392° F. (473 K) were used to model the heat transfer
bum duration were fixed user inputs. The 50% burn point
between the combustion gas and in-cylinder components forprovides a more direct means of phasing the combustion 50 the remaining studies.
event, as there is no need to track spark timing and ignition
The engine friction was characterized within GT-Power
delay. The 10 to 90% burn duration is the crank angle interval
using the Chen-Flynn correlation, which is an experimentrequired to burn the bulk of the charge, and is the common
based empirical relationship relating cylinder pressure andterm for defining the duration of the combustion event. The
mean piston speed to total engine friction. The coefficients
output of the Wiebe combustion model is a realistic non- 55 used in the Chen-Flynn correlation were adjusted to give
instantaneous heat release curve, which is then used to calrealistic friction values across the speed range.
culate cylinder pressure as a function of crank angle (° CA).
2.3 Summary of Results of the Conventional Engine
The Wiebe function is an industry standard for an empirical
Table 3 summarizes the performance results for the twoheat release correlation, meaning that it is based on previous
cylinder conventional four-stroke engine model. The results
history of typical heat release profiles. It provides an equa- 60 are listed in terms of indicated torque, indicated power, indition, based on a few user-input terms, which can be easilycated mean effective pressure (IMEP), indicated thermal effiscaled and phased to provide a reasonable heat release profile.
ciency (ITE), pumping mean effective pressure (PMEP), fricFIG. 13 shows a typical Wiebe heat release curve withtion mean effective pressure (FMEP), brake torque, brakesome of the key parameters denoted. As shown, the tails of thepower, brake mean effective pressure (BMEP), brake thermalheat release profile (<10% burn and >90% bum) are quite 65 efficiency (BTE), volumetric efficiency, and peak cylinderpressure. For reference, mean effective pressure is defined aslong, but do not have a strong effect on performance due to thesmall amount of heat released. At the same time, the actual
the work per cycle divided by the volume displaced per cycle.

US 8,006,656 B2Parameter

19

20

TABLE3eters were the TDC phasing and compression and Expansion
Ratios. The modified engine parameters were summarized inTables 4 and 5

1400 rpm

1800 rpm

2400 rpm

3000 rpm

Summary of Predicted Conventional Engine Performance(English Units)

TABLE4

93.4
Indicated Torque (ft-lb)
90.6
92.4
90.7
Indicated Power (hp)
24.2
31.7
42.7
51.8
Net IMEP (psi)
135.9
138.5
140.1
136.1
38.0
ITE (%)
37.5
37.9
38.2
-0.6
-1.2
-2.4
-4.0
PMEP (psi)FMEP (psi)
15.5
17.5
20.5
23.5
80.3
80.7
79.7
75.1
Brake Torque (ft-lb)
21.4
27.7
36.4
42.9
Brake Power (hp)
119.6
112.6
BMEP (psi)
120.4
121.0
BTE (%)
33.2
33.1
32.6
31.5
89.0
89.5
87.2
Vol. Eff. (%)
88.4
Peak Cylinder Pressure
595
600
605
592
(psi)
Summary of Predicted Conventional Engine Performance (SI Units)
Indicated Torque (N-m)
Indicated Power (kW)Net IMEP (Bar)ITE(%)PMEP (bar)FMEP (Bar)
Brake Torque (N-m)
Brake Power (Kw)
BMEP (bar)BTE(%)
Vol. Eff. (%)
Peak Cylinder Pressure
(bar)

122.9
18.0
9.4
37.5
-0.04
1.07
108.9
16.0
8.3
33.2
88.4
41.0

125.2
23.6
9.6
37.9
-0.08
1.21
109.4
20.6
8.3
33.1
89.0
41.4

126.7
31.8
9.7
38.2
-0.17
1.42
108.1
27.2
8.2
32.6
89.5
41.74

123.0
38.6
9.4
38.0
-0.28
1.62
101.8
32.0
7.8
31.5
87.2
40.8

Split-Cycle Engine Parameters (Compression Cylinder)

10

15

Parameter

Value

Bore
StrokeConnecting Rod Length
Crank Throw
Displacement Volume
Clearance Volume
Compression Ratio
Cylinder Offset
TDC Phasing
Engine Speed
Air/Fuel Ratio

4.410 in (112.0 mm)
4.023 in (102.2 mm)
9.6 in (243.8 mm)
2.011 in (51.1 mm)
61.447 in3 (1.007 L)
0.621 in 3 (O.GlO L)
100:1
1.00 in (25.4 mm)
25°CA
1400 rpm
18:1

20

TABLES
Split-Cycle Engine Parameters (Expansion Cylinder)
25

30

Parameter

Value

Bore
StrokeConnecting Rod Length
Crank Throw
Displacement Volume
Clearance Volume
Expansion Ratio
Cylinder Offset

4.000 in (101.6 mm)
5.557 in (141.1 mm)
9.25 in (235.0 mm)
2.75 in (70.0 mm)
69.831 in3 (1.144 L)
0.587 in 3 (O.GlO L)
120:1
1.15 in (29.2 mm)

Referring to FIG. 14 performance is plotted in terms ofbrake torque, brake power, BMEP, volumetric efficiency,
FMEP, and brake thermal efficiency across the speed range.
The valve events were initially set using measured lift profiles
Referring to FIGS. 15A and B, the GT-Power GUI for thefrom an existing engine. The timing and duration of the intake 35 split-cycle engine model is shown. Intake air flows from theambient source into the intake manifold, represented by pipeand exhaust valves events were tuned to yield representative
intk-bypass and junction intk-splitter. From there, the intake
volumetric efficiency values across the speed range. As
air enters the intake ports (intportl, intport2) where fuel isshown in FIG. 14, the volumetric efficiency is approximately
injected and mixed with the airstream. At the appropriate time
90% across the speed range, but began to drop off slightly at
3000 rpm. Similarly, the brake torque values were fairly flat 40 of the cycle, the intake valves (vil-y) open while the piston incylinder camp is on its downstroke (intake stroke). The air
across the speed range, but tailed off slightly at 3000 rpm. The
and fuel mixture are admitted into the cylinder during thisshape of the torque curve resulted in a near linear powerstroke, after which time the intake valves close. After the
curve. The trend of brake thermal efficiency across the speed
intake stroke, the piston rises and compresses the mixture torange was fairly consistent. There was a range of 1. 7 points of
45 a high temperature and pressure. Near the end of the comthermal efficiency from the maximum at 1400 rpm of33.2%
pression stroke, the pressure is sufficient to open the checkto the minimum at 3000 rpm of 31.5%.
valve (check) and push air/fuel mixture into the crossover
passage. At this same time, the power cylinder has just com3.0 Split-Cycle Engine Model
pleted the exhaust stroke and passed TDC. At approximately
50 this time, the crossover valve (cross valve) opens and admits
A model of the split-cycle concept was created in GTair from the crossover passage and from the camp cylinder,
Power based on the engine parameters provided by the Scuwhose piston is approaching TDC. At approximately the timederi Group, LLC. The geometric parameters of the compresof the camp cylinder's piston TDC (i.e. after power cylinder'ssion and expansion cylinders were unique from one anotherpiston TDC by the phase angle offset), the crossover valve
and quite a bit different from the conventional engine. The 55 closes and the spark plug is energized in the power cylinder.
validity of comparison against the conventional engine
The mixture bums, further raising the temperature and presresults was maintained by matching the trapped mass of thesure of the mixture and pushing down on the power piston
intake charge. That is, the split-cycle engine was made to havethrough the expansion or power stroke. Near the end of thethe same mass trapped in the compression cylinder after
expansion stroke, the exhaust valve opens and the piston
intake valve closure as the conventional; this was the basis of 60 begins to rise, pushing the exhaust out of the cylinder via thethe comparison. Typically, equivalent displacement volume isexhaust valves (vel, ve2) into the exhaust ports (exhportl,
used to insure a fair comparison between engines, but it is
exhport2). Note that the compression and exhaust strokes asvery difficult to define the displacement of the split-cycle
well as the intake and power strokes are taking place atroughly the same time but on different cylinders. From theengine; thus equivalent trapped mass was used as the basis.
65 exhaust ports, the exhaust is transmitted into the exhaust
3.1 Initial Split-Cycle Model
Several modifications were made to the split-cycle engine
manifold (exh-jcn) and from there to the end enviroument
model. It was found that some of the most significant param(exhaust) representing the ambient.

US 8,006,656 B2

21

22Note that the layout of the model is very similar to the
TABLE 6-continued
conventional engine model. The intake and exhaust ports andSplit-Cycle Engine Breathing and Combustion Parameters
valves, as well as the multi-port fuel injectors, were taken
directly from the conventional engine model. The crossover
All referenced topassage was modeled as a curved constant diameter pipe with
TDC of power
Value
Parameter
cylinderone check valve at the inlet and poppet valves at the exit. In theinitial configuration, the crossover passage was 1.024 in (26.0134° ATDC (power) 134° ATDC
Exhaust Valve Opening (EVO)mm) diameter, withfour0.512 in (13.0 mm)valves at the exit.
358° ATDC
2° BTDC (power)
Exhaust Valve Closing (EVC)
The poppet valves feeding the expansion cylinder were0.362 in (9.18 mm)
Peak Exhaust Valve Liftso BTDC (power) 355° ATDC
10 Crossover Valve Openingreferred to as the crossover valves.
(XVO)
Though the crossover passage was modeled as a curved
Crossover Valve Closing (XVC) 25° ATDC (power)
constant diameter pipe having a check valve inlet and poppet
Peak Crossover Valve Lift
0.089 in (2.27 mm)
valve outlet, one skilled in the art would recognize that other
50% Bum Point
37° ATDC (power)
Combustion Duration (10-90%) 24°CA
configurations of the above are within the scope of this invention. For example, the crossover passage may include a fuel 15
injection system, or the inlet valve may be a poppet valve
Additionally, FIG. 17 provides a graph of the compression
rather than a check valve. Moreover various well known
and expansion piston positions, and valve events for the splitvariable valve timing systems may be utilized on either of thecycle engine.
crossover valve or the inlet valve to the crossover passage.
One of the first steps was to check the clearance between
Referring to FIG. 16, a model of the split-cycle engine was 20
the crossover valve and power cylinder piston. The crossover
constructed using an MSC.ADAMS® dynamic analysis softvalve is open when the expansion cylinder piston is at TDC,
ware package to confirm the piston motion profiles and proand the piston-to-head clearance is 0.040 in (1.0 mm). Thereduce an animation of the mechanism. MSC.ADAMS® softwas
interference indicating valve-to-piston contact. Attempts
ware, owned by MSC.Software Corporation of Santa Ana,
Calif., is one of the most widely used dynamics simulation 25 were made to fix the problem by adjusting the phasing of thecrossover valve, but this resulted in a 1 to 2 point penalty insoftware packages in the engine industry. It is used to calcuindicated thermal efficiency (ITE) across the speed range.
late forces and vibrations associated with moving parts in
The trade-offs were discussed and it was decided that it would
general. One application is to generate motions, velocities,
be better to alleviate the interference and return to the previand inertial forces and vibrations in engine systems. FIG. 16
shows a schematic representation of the MSC.ADAMS® 30 ous phasing, thus retaining the higher ITE values. Possiblemodel.
solutions to be considered include valve reliefs in the pistoncrown, recessing the valves in the cylinder head, or outward
Once the split-cycle engine model was producing positive
opening valves.
work, there were several other refinements made. The timingof the intake valve opening (IVO) and exhaust valve closing
Next, the number of crossover valves was reduced from
(EVC) events were adjusted to find the best trade-offbetween 35 four to two, with the valves sized to match the cross-sectional
valve timing and clearance volume as limited by valve-toarea of the crossover passage outlet. For the 1.024 in (26. mm)
position interference. These events were investigated during
diameter crossover passage outlet, this resulted in two 0.724
the initial split-cycle modeling efforts and optimum IVO andin (18.4 mm) valves as compared to four 0.512 in (13.0 mm)
EVC timings were set. IVO was retarded slightly to allow forvalves. This change was made to simplifY the crossover valve
the compression piston to receive some expansion work from 40 mechanism and make the expansion side cylinder head more
the high gas pressure remaining after feeding the crossover
like a typical cylinder head with two intake valves.
passage. This precluded the trade-offbetween reducing clearThe wall temperature solver in GT-Power was used toance volume and early IVO for improved breathing. The
predict the piston, cylinder head, and cylinder liner wall temengine breathed well, and the late IVO allowed the piston toperatures for both the conventional and split-cycle engines.
45 Originally, it was assumed that aluminum pistons would berecover a bit of expansion work.
used for both the conventional and split-cycle engines. The
EVC was advanced to produce a slight pressure build-up
predicted piston temperatures for both the conventional
prior to crossover valve opening (XVO). This helped reduceengine and split-cycle compression cylinder piston were well
the irreversible loss from dumping the high-pressure gas fromwithin standards limits, but the split-cycle power cylinder
the crossover chamber into a large volume low-pressure resso piston was approximately 266° F. (130° C.) over the limit. To
ervmr.
address this concern, the power cylinder piston was changed
The Wiebe combustion model was used to calculate theto a one-piece steel oil-cooled piston. This brought the averheat release for the split-cycle engine. Table 6 summarizes theage temperature to within the limit for steel-crown pistons.
valve events and combustion parameters, referenced to TDC
The average cylinder wall temperature for the split-cycleof the expansion piston, with the exception of the intake valve
events, which are referenced to TDC of the compression 55 power cylinder was approximately 140° F. (60° C.) higher
piston.
than the conventional engine. This could lead to problems
with lube oil retention. The wall temperatures were calculated
across the speed range and then averaged and applied as fixed
TABLE6wall temperatures for all remaining studies. Fixed surface
Split-Cycle Engine Breathing and Combustion Parameters
60 temperatures for the expansion cylinder components were
860° F. (733 K) for the piston, 629° F. (605K) for the cylinder
All referenced tohead, and 552° F. (562K) for the liner. For the compression
TDC of powerValuecylinderParameter
cylinder components, the surface temperatures were 399° F.
(473K) for the piston, 293° F. (418K) for the cylinder head,
17° ATDC (camp)
42° ATDC
Intake Valve Opening (NO)65 and 314 o F. (430K) for the liner.
Intake Valve Closing (IVC)
174°BTDC(comp) 211°ATDC
Peak Intake Valve Lift
0.412 in (10.47 mm)Table 7 summarizes the performance results for the initial
split-cycle engine model. The results are listed in terms ofUS 8,006,656 B2
23
indicated torque, indicated power, indicated mean effective
pressure (IMEP), indicated thermal efficiency (ITE), andpeak cylinder pressure.

24

gains in thermal efficiency across the speed range as a resultof optimizing crossover passage diameter were minimal (lessthan 0.3 points ITE).
3.2.2 TDC Phasing
TABLE 7
Sweeping the TDC phasing between the compression andpower
cylinders exerted a significant influence on thermal
Parameter
1400 rpm 1800 rpm 2400 rpm 3000 rpm
efficiency. The TDC phasing was swept between 18° and 30°
Summary of Predicted Engine Performance (English Units)
CA. At each step, the 50% bum point and crossover valve
timingwere adjusted to maintain the phasing such that the
Indicated Torque (ft-lb)
92.9
80.8
91.9
88.1
10
Indicated Power (hp)
24.8
31.5
40.3
46.2
10% bum point occurred at or after the crossover valve closNet IMEP (psi)
53.8
53.2
51.0
46.8
ing (XVC) event. This was intended to prevent flame propaITE(%)
36.1
35.8
34.6
33.0
630
656
730
807
Peak Cylinder Pressure,gation into the crossover passage. The most promising conCompression Cylinder
figuration came from a TDC phasing of 20° CA. This(psi)
15 demonstrated moderate gains across the speed range (1.3 toPeak Cylinder Pressure,
592
603
623
630
Expansion Cylinder (psi)1.9 points ITE relative to the previous 25° TDC phasing).
Summary of Predicted Engine Performance (SI Units)
Further studies to optimize the crossover valve duration andlift resulted in minimal improvement (less than 0.2 points
Indicated Torque (N-m)
126.0
124.6
119.4
109.6
Indicated Power (kW)
18.5
23.5
30.0
34.4
20 ITE).
Net IMEP (bar)
3.71
3.67
3.52
3.23
3.2.3 Combustion Duration
ITE(%)
36.1
35.8
34.6
33.0
Changing the combustion duration, or 10 to 90% bum
43.4
45.2
50.3
55.6
Peak Cylinder Pressure,Compression Cylinderrates, also exerted a strong influence on the thermal efficiency.
(bar)
The initial setting for 10 to 90% combustion duration was set
Peak Cylinder Pressure,
40.9
41.6
43.0
43.5
25 at 24° CA, which is a rapid burn duration for typical SI
Expansion Cylinder (bar)engines. The most important objective was to maintain thesame type of combustion duration between the conventional
FIG. 18 plots the performance in terms of indicated torque,and split-cycle engines. However, due to theories relating toindicated power, and new IMEP across the speed range. The
faster burn rates that might be inherent in the split cycle
trend of indicated torque and net IMEP is flat at 1400 and 30
engine, the engine's sensitivity with regards to a faster com1800 rpm, but drops off at the higher speeds. The power curvebustion event was examined. Reducing the 10 to 90% bum
is somewhat linear. Most of the emphasis was focused on
duration (increasing the bum rate) from 24 o CA to 16° CA
tuning for the 1400 rpm operating point, thus there was notshowed gains of up to 3 points ITE across the speed range.
much effort expended in optimizing high-speed engine
This study was repeated for the conventional engine model
35
operation.
to establish a reference point for comparison. The gains for
3.2 Parametric Sweeps
the conventional engine were limited to 0.5 point ITE. For theParametric sweeps were conducted to determine the influconventional engine, combustion takes place at a near conence of the following key variables on indicated thermal
stant volume.
efficiency:
Referring to FIG. 19, the log pressure vs.log volume (log40
log P-V) diagram for the conventional engine at the 24 o CA
Crossover passage diameter,
10 to 90% burn duration is shown. When compared to the
Crossover valve diameter,
ideal Otto cycle constant volume heat addition line, there is a
TDC phasing,shaded region above where the combustion event transitions
Crossover valve timing, duration, and lift,
45 into the expansion stroke. By decreasing the burn duration to
10 to 90% burn duration,
16° CA, there is an increase in the amount of fuel burned near
Bore-to-Stroke ratio (constant displacement)
TDC
that results in increased expansion work. In other words,
Expansion cylinder Expansion Ratio,
theshadedregion gets smaller, and the P-V curve more
Heat transfer in crossover passage, andclosely approximates the ideal Otto cycle. This leads to slight
In-cylinder heat transfer for expansion cylinder.
50 improvement in thermal efficiency. Engine manufacturers
For all the parametric sweeps conducted, several runs were
have invested significant development efforts in optimizing
conducted at the 1400 rpm engine speed condition to deterthis trade-off for incremental improvements.
mine the most promising configuration. Once that configuraReferring to FIG. 20, the pressure volume diagram for thetion was identified, runs were conducted across the speed
split-cycle engine is shown. The split-cycle engine expansion
range. The results are presented in terms of gains or losses in 55 cylinder undergoes a much larger change in volume during
ITE relative to the results from the initial split-cycle engine
the combustion event when compared to the conventionalmodel or previous best case.
engine. This is illustrated in FIG. 20. The black line represents
3.2.1 Crossover Passage Diameterthe 24 o CA to 10 to 90% burn duration.
The crossover passage diameter was varied from 0.59 in
Thermal efficiency increases as combustion is shifted
60(15.0 mm) to 1.97 in (50.0 mm). At each step, the crossover
towards TDC for the split-cycle engine, but advance of thevalve diameter was changed such that the area of the two
10% bum point is limited by the timing of the crossover
valves matched the area of the crossover passage outlet. The
closing (XVC) event. Reducing the 10 to 90% burn duration
effectively advances combustion, resulting in more pressure
most promising configuration for the crossover passage was
1.18 in (30 mm) diameter inlet and outlet cross sections with 65 acting over a reduced change in volume. Thus, reducing thebum duration yields larger gains with the split-cycle engine
two 0.83 in (21.2 mm) crossover valves. The inlet was modthan with the conventional engine.
eled with a check valve with a realistic time constant. The

US 8,006,656 B2

25

26

A typical 10 to 90% burn duration or a conventional spark
TABLES
ignited gasoline engine is between 20° and 40° CA. One ofIndicated Thermal Efficiency Predictions for Split-Cycle Engine
the limiting factors in increasing bum rates is how muchturbulence can be produced inside the cylinder, thus wrin1800 2400
3000
Configuration
1400 rpm rpm
rpm
rpm
kling the flame front and speeding up the flame propagation
across the cylinder. The GT-Power Wiebe combustion model
Conventional engine model
37.5
27.9 38.2
38.0
Initial split-cycle engine model
36.1
35.8 34.6
33.0
does not account for this level of complexity. It was hypoth30-mm crossover passage
36.2
36.0 34.9
33.3
esized that, due to the intense motion and late timing of the
20° TDC phasing
37.5
37.5 36.6
35.2
crossover flow, the split-cycle engine expansion cylinder may 10 16° 10 to 90% burn duration
40.6
40.6 40.0
38.6
1.5-mm ceramic coating (crossover)
41.3
41.4 40.9
39.6
experience a much larger degree of bulk air motion and turExpansion cylinder ceramic
42.8
42.9 42.6
41.5
bulence at the time of combustion, thus leading to higher
components
flame speeds than the conventional engine. It was decided topursue computational fluid dynamics (CFD) analysis to more
Referring to FIG. 21, these results are displayed graphiaccurately model the combustion event and determine the 15
cally. As a basis of comparison, the conventional engine
types of bum rates possible for the split-cycle engine. This
yielded indicated thermal efficiencies in the range of37 .5% totopic is covered in Section 3.3.
38.2% at similar power levels as the split-cycle engine.
3.2.4 In-Cylinder Geometry
Speeding up the bum rates had the most significant influence
In the next set of parametric studies, the in-cylinder geom- 20 of any of the variables investigated. The increased bum rates
etry was varied to determine the influence on thermal effiallowed the thermal efficiencies of the split-cycle engine tociency. The bore-to-stroke ratio was varied independently forrise above the levels predicted for the conventional engine by
the compression and power cylinders, holding displacement
approximately 3 points. Further potential increases wereconstant for each. For the compression cylinder, the bore-todemonstrated with the use of ceramic coatings.
stroke ratio was swept from 0.80 to 1.20. The most promising 25 3.3 Combustion Analysis
compression cylinder bore-to-stroke ratio for the 1400 rpm
The parametric sweep conducted in GT-Power demonengine speed was 0.90 (0.3 point ITE gain). However, thisstrated that the 10 to 90% bum duration had a significant
value did not result in gains for the other engine speeds. The
influence on the ITE of the split-cycle engine. It was also
decrease in bore-to-stroke ratio translates to a longer stroke
hypothesized that the split-cycle engine expansion cylinderand connecting rod, which increases engine weight, particu- 30 may experience higher levels of in cylinder bulk air motion
larly for the engine block. There were no gains demonstrated
and turbulence as compared to the conventional engine, thusfrom changing the bore-to-stroke ratio of the expansion cylyielding faster burn rates. The Wiebe combustion model usedinder. Increasing the Expansion Ratio of the expansion cylduring the GT-Power cycle simulation studies produces heatinder from 120 to 130 showed a gain of0.7 point ITE for the
release curves based on user inputs for the 50% burn point and1400 rpm operating point. There was a slight penalty in ITE at 35 10 to 90% bum duration. It provides a rough approximation of
the higher engine speeds, however. All signs indicate that if
the combustion event, but does not account for the effects ofthe engine were tuned for a 1400 rpm application, there would
increased turbulence.
Computational fluid dynamics (CFD) was utilized to test
be some benefit in ITE from changing the compression cylthe hypothesis and quantifY the 10 to 90% bum durationinder bore-to-stroke ratio and the power cylinder Expansion
Ratio. However, if tuning across the speed range, the values 40 achievable with the split-cycle engine concept. Computashould be left unchanged.
tional Fluid Dynamics refers to a field of software that
3.2.5 Heat Transfer
reduces a complex geometric field into tiny pieces (referred toas a "elements" which are separated by the "grid"). The
Ceramic coatings were modeled and applied to the crossapplicable governing equations (fluid flow, conservation ofover passage to quantify potential gains in thermal efficiency
due to retained heat and increased pressures in the passage. 45 mass, momentum, energy) are then solved in each of theseUsing a thermal conductivity of 6.2 W/m-K, the emissivity
elements. Stepping forward in time and completing theseand coating thickness were varied. The wall thickness, whichcalculations for each element for each time step allows thewas varied from 0.059 in (1.5 mm) to 0.276 in (7 mm), did notsolving of very complex flow fields but requires high computational power.
exert much influence on thermal efficiency. The 0.059 in (1.5mm) thickness is a typical value used for ceramic coatings of 50
CFD models were constructed ofboth the conventional andengine components, so it was used as the default. Varying the
split-cycle engines to provide comparative analyses. The
emissivity, which can vary anywhere from 0.5 to 0.8 for aintake valve events and spark timing were adjusted for theceramic material, led to a shift of 0.2 points ITE, with theconventional engine to match the trapped mass and 50% bum
lower value of 0.5 yielding the best results. With this emispoint from the cycle simulation results. The resulting 10 tosivity, there was a predicted gain of0.7 points ITE across the 55 90% burn duration from CFD was approximately 24 o CA,
which matched the value used in the GT-Power Wiebe comspeed range.
There was no quick straight forward method in GT-Powerbustion model.
for applying ceramic coatings to the in-cylinder components.
For the split-cycle model, the inputs included fixed wall
Rather than invest a great deal of time creating a sub-model totemperatures assuming ceramic coating on the crossover pasperform the necessary calculations, the material properties 60 sage, but no ceramic components in the expansion cylinder.
for the power cylinder piston and cylinder head were
The early portion of the bum occurs with the crossover valve
switched to ceramic. The results suggest that there could be
open. The interaction between the intake charge from thecrossover passage and the expansion cylinder pressure risegains as high as 2 points ITE across the speed range fromfrom combustion effects the trapped mass. Several iterations
using the ceramic components.
3.2.6 Summary of Results ofiTE on the Split-Cycle Engine 65 were required to match the trapped mass from the convenTable 8 below tracks the changes in ITE through the course
tional engine within 4%. The first set of results had a signifiof the parametric studies.
cant amount of overlap with approximately 35% of the total

US 8,006,656 B2

27

28

combustion event (i.e. from the 0% point to the 100% point ofare easier to oxidize under lean operating conditions thanNOx using readily-available exhaust after treatment devices
combustion) occurring prior to crossover valve closing. (Thissuch as oxidation catalysts.
will be referred to as 35% "bum overlap" from hereon.) The
Referring to FIG. 24, the predicted NOx emissions for all
CFD model had combustion disabled in the crossover pasthree cases, i.e. conventional engine, split-early (5% bum
sage. However, by reviewing the results, it became clear thatoverlap) and split-late (35% bum overlap), are shown. Expethis amount of overlap would have more than likely resulted
rience indicates that the relative NOx trend between cases isin flame propagation into the crossover passage. The resulting
accurately predicted, but that the absolute magnitude may not
10 to 90% bum duration was approximately 10° CA.
be. Both of the split-cycle cases have combustion events later
Referring to FIG. 22, the case with the 35% burn overlap is
10 in the cycle than the conventional case, resulting in less overillustrated as calculated via the CFD analysis. The crossover
all time at high temperatures, and thus less NOx than thevalve 250 is closed after approximately 35% of the burnconventional case. The later timing case produced very little
occurs and the expansion piston 252 is being driven downNOx because the late combustion resulted in lower cylinder
ward by the hot gases. The flame front 254 (the dark shaded
temperatures. The expansion cycle was well underway whenarea) has progressed passed the crossover valve seat 256. 15 combustion was occurring.
Accordingly, it is likely that in this embodiment the flame
The lower cylinder temperatures for the late bum splitfront 254 would be able to creep into the crossover passage
cycle case resulted in increased CO emissions when com258.
pared to both the conventional engine and the early timing
Another iteration was conducted to reduce the bum oversplit cycle engine case. The final CO concentrations were 39,
lap. The target was less than 10% of the burn occurring prior 20 29, and 109 ppm for the conventional, early timing splitto crossover valve closing. Again, several iterations werecycle, and late timing split cycle respectively.
required to match the trapped mass. This case resulted in
3.4 Friction Study
The friction model used in GT-Power is based on the Chenapproximately 5% of the total combustion event (i.e. from the
Flynn correlation, which predicts friction using the following
0% point to the 100% point of combustion) occurring prior tocrossover valve closing. The 10 to 90% bum duration was 25 empirical relationship:
approximately 22° CA. The amount of overlap between the
crossover valve and combustion events exerted a significant
FMEP: friction mean effective pressure (or friction torqueinfluence on the burn duration.
per displacement).
Referring to FIG. 23, the case of the 5% burn overlap is
a,b,c,d: correlation coefficients (tuning parameters)
illustrated as calculated via the CFD analysis. The crossover 30
PCP: peak cylinder pressure, andvalve 250 is closed after approximately 5% of the bum occurs
VP: mean piston speed.
and the expansion piston 252 is being driven downward by theThis correlation has been well developed over some time
hot gases. The flame front 254 (the dark shaded area) has notfor conventional piston engines and reasonable values for the
progressed past the crossover valve seat 256. Accordingly, it 35 correlation coefficients have been validated against experiis likely that in this embodiment the flame front 254 would notmental data. However, the empirical mode does not take into
be able to creep into the crossover passage 258.
account the unique piston motion and connecting rod angle of
One interesting discovery from the CFD analysis was thatthe split-cycle engine concept.
the split-cycle engine appears to have a potential inherent
The dominant source of engine rubbing friction comes
advantage over the conventional engine in terms ofNOxemis- 40 from the piston assembly. More specifically, the dominantsions. The predicted NOx emissions for the 10° CA 10 to 90%
source of piston assembly friction comes from contact
bum duration split-cycle engine case were roughly 50% of the
between the piston rings and cylinder liner. To determine theinherent differences in engine friction between the convenNOx emissions predicted for the conventional engine, whilethe 22° CA 10to 90% bumdurationcaseresultedin approxitional and split-cycle engines, friction calculations were permately 20% of the conventional engine NOx emissions. The 45 formed outside of GT-Power. Piston thrust loading was calculated as a function of the cylinder pressure vs. crank angle
high rate of expansion during combustion found in the splitcycle engine will result in a reduction of the maximum enddata imported from GT-Power in a spreadsheet format. Fricgas temperatures that are normally experienced in a convention force was determined by multiplying this force by anaverage (constant) coefficient of friction value. The friction
tional engine, which burns at almost constant volume.
Therefore the trend of these results appears to be reasonable. 50 work was calculated by integrating the F -dx work throughout
the stroke in increments of0.20 CA. It was assumed that the
Typical SI gasoline automotive engines operate at stoichiosum of F-dx friction work accounted for half of the total
metric or slightly rich air/fuel ratios at full load. Thermal
engine friction. The average coefficient of friction value was
efficiency tends to improve with lean air/fuel ratios, but withincreased NOx emissions and severely degraded catalyst perdetermined by matching the predicted friction work from theformance. The inability of the catalyst to effectively reduce 55 spread sheet to friction work predicted from the Chen-Flynn
NOx emissions under these conditions further aggravates thecorrelation for the conventional engine at 1400 rpm. This
tailpipe NOx levels. The predicted NOx emissions for thevalue was then applied to the split-cycle engine to predict theconventional engine operating at 18: 1 air/fuel ratio are likely
piston assembly friction. The remaining half of friction was
higher than what would be representative of typical engines
assumed to remain constant between the two engine configu60 rations, as it deals with valve train, bearing friction, andoperating at stoichiometric or slightly rich air/fuel ratios.
accessory losses. FMEP varies with engine speed, and theThese results have not been correlated to experimental data
1400 rpm point was selected to remain consistent with theand emissions predictions from numerical models tend to behighly dependent on tracking of trace species through theprevious parametric studies.
combustion event. If these results were confirmed on an
The amount of friction work accounts for the differences
actual test engine, they would constitute a significant advan- 65 between indicated and brake work for a given engine. The
tage of the split-cycle engine concept. Predicted CO emisfriction torque and power values were very similar between
sions were higher for the split-cycle engine, but these species
the conventional and the split-cycle engines with 22° 10 toUS 8,006,656 B2

29

30

The 1oo CA 10 to 90% burn duration from run# 181 yielded
90% burn duration. However, the results suggest that thea gain of approximately 5.0 points BTE over the conventional
split-cycle engine may have a slightly higher mechanicalefficiency than the conventional engine as the 10 to 90% burnengine. However, in the current configuration, these condiduration is shortened from 22° CA. For example, at the 16°
tions would likely lead to flame propagation into the crossover passage. The 22° CA 10 to 90% burn duration from rnn
CA 10 to 90% burn duration, the split-cycle engine had a 1.0
point advantage in mechanical efficiency, which translates to#183 is realistically achievable with respect to avoidance of
a 1.0 point gain in BTE.
flame propagation into the crossover passage, and resulted in
Referring to FIG. 25, the reasons for this trend is illusa gain of approximately 0.7 points ITE.
trated. FIG. 25 plots the expansion piston thrust load versus
3.6 Investigation Of Lower Limits for Significant Parameters
crank angle, referenced to TDC of the expansion piston, for 10
The studies conducted during construction of the initial
the 10° CA and 22° CA 10 to 90% burn duration cases. The
split-cycle model and subsequent parametric sweeps identi10° CA 10 to 90% burn duration resulted in a mechanicalfied Compression Ratio, Expansion Ratio, TDC phasing, andefficiency approximately 1.2 points higher than the 22° CA
burn duration as significant variables affecting engine perforcase. For the 10° CA 10 to 90% burn duration case, thrust 15 mance and efficiency. Additional cycle simulation rnns wereloading increased more rapidly after the connecting rod
performed to identify lower limits of Compression Ratio,
passed through the oo angle point. Even though the 10° CA
Expansion Ratio, TDC phasing, and crossover valve lift andcase reached a higher peak thrust load, the 22° CA caseduration where engine performance and/or efficiency tails
maintained a slightly higher thrust load than the 1oo CA caseoff.
through the remainder of the stroke. When the integration of
The baseline for comparison was the split-cycle engine
20
F -dx is performed, the 10° CA had lower piston friction work.
with a 10 to 90% burn duration of 22° CA (Rnn #183).
3.5 Summary of the Results for the Split-Cycle Engine
Sweeps were conducted from this base configuration to quanThe resulting burn rates from the CFD combustion analysis
tify indicated power and ITE as functions of Compression
were used to set up and run additional iterations in GT-Power
Ratio, Expansion Ratio, TDC phasing, and crossover valve
for the split-cycle engine. Table 9 summarizes the results andcompares them to the conventional engine in terms of indi- 25 lift and duration. It should be noted that the inter-dependent
effects of these variables exert a significant influence on thecated, friction, and brake values. All runs were conducted at
performance and efficiency of the split-cycle engine concept.
an engine speed of 1400 rpm.
For this study, the effects of each of these variables wereisolated. No sweeps were conducted to analyze the combined
TABLE930 influence of the variables. Altering each of these variables
exerts a strong influence on trapped mass, so relative comSplitSplitSplitCycle
Cycle
Cycle
parisons to run #183 or the conventional engine may not be(Run
Conventional
(Run
(Run
valid.
Parameter(Run #96)
#180)
#181)
#183)
FIG. 26 shows the indicated power and ITE for various
Summary of Results (English Units)
35 Compression Ratios. The baseline was set at a Compression
Ratio of 100:1. Reducing this value to 80: 1 results in a 6%
10-90% Burn Duration
22
10
24
16
in airflow and indicated power. ITE decreases withdecrease(eCA)
Compression Ratio also, but more dramatically at 40:1 and
50% Bum Point (eATDC)
24
32
28
10
93.7
Indicated Torque (ft-lb)
91.8
102.4
103.6
lower.
24.6
Indicated Power (hp)
24.2
27.0
27.2
40
FIG. 27 plots indicated power and ITE for various Expan38.2
ITE(%)
37.5
41.2
42.7
sionRatios.
Indicated power was somewhat steady with slight
10.4
10.4
10.5
Friction Torque (ft-lb)
10.3
increases in airflow as Expansion Ratio was decreased from
2.78
Friction Power (hp)
2.76
2.79
2.74
83.3
Brake Torque (ft-lb)
81.4
92.0
93.3
the initial value of 120:1. At 40:1, airflow into the cylinder
22.3
Brake Power (hp)
21.4
24.5
24.9
was 5% high with a moderate drop in ITE. At 20:1, airflow
88.9
Mechanical Efficiency (%)
88.7
89.8
90.1
45 was 9% high, indicated power was 4% low, and ITE was more
33.9
38.4
BTE(%)
33.2
37.0
than 4.0 points lower than the baseline.
Summary of Results (SI Units)
FIG. 28 plots the same data for various TDC phase angles.
10-90% Burn Duration
22
10
24
16
During these rnns, the phasing for the crossover valve and(eCA)
combustion events were left unchanged in relation to the
50% Bum Point (eATDC)
32
24
28
10
50 expansion piston's TDC. There was a moderate drop in ITE as
127.0
Indicated Torque (N-m)
124.4
138.9
140.5
18.4
Indicated Power (kW)
18.0
20.2
20.3
the TDC phasing was reduced from the original value of 20°
38.2
42.7
ITE(%)
37.5
41.2
CA. Airflow and indicated power decrease more sharply with
14.1
13.9
Friction Torque (N-m)
14.1
14.2
TDC phase angle. Also, friction is increased due to higher
2.07
2.04
2.07
2.08
Friction Power (kW)peak cylinder pressures. At a TDC phasing ofl oo, airflow and
112.9
126.5
Brake Torque (N-m)
110.3
124.7
16.6
18.6
16.0
18.3
Brake Power (kW)
55 indicated power were approximately 4% down from the base88.9
90.1
Mechanical Efficiency (%)
88.7
89.8
line, with a 0.7 point drop in ITE, as well as an additional 0.5
33.9
38.4
BTE(%)
33.2
37.0
point penalty in BTE due to increased friction.
The leveling out of performance at higher phasing offset
Split-cycle rnn #180 represents the 16° CA 10 to 90% burnangles may not be representative of realistic engine operation.
duration from the previous parametric sweeps. Run #181 60 At this point, with the approach taken here in the investigation
oflower limits section of the study, the crossover valve event
represents the first iteration ofCFD combustion analysis conducted on the split-cycle engine model. This run resulted inand compression event are grossly mis-timed such that the
split-cycle concept is not accurately represented. At the late
approximately 35% of the burn occurring prior to crossover
valve closing, which would likely lead to flame propagation
phasing, the crossover valve opens before the compressor
into the crossover passage. Run #183 represents the second 65 cylinder begins charging the crossover in earnest, such thatthe basic process is to accumulate mass in the crossover
iteration of CFD combustion analysis, with approximately
5% of the burn occurring at crossover valve closing.
passage on one cycle and then allow it to enter the powerUS 8,006,656 B2
31

32

cylinder on the next cycle. That is the reason for the flatness ofthe phase angle should be less than or equal to 50 degrees
and was set at approximately 20 degrees for this study;
the curve at those high phasing angles.
andFIG. 29 plots the same results as a function of crossoverthe crossover valve duration should be less than or equal tovalve duration and lift. Comparing tables 2 and 6, it can be
69 degrees and was set at approximately 25 degrees forseen that the crossover valve duration of the split-cycle engine
this Study.
(i.e., 30° CA) is much smaller than the intake and exhaust
Moreover, the crossover valve duration and the combustion
valve durations of the conventional engine (225° CA and 270°
duration should overlap by a predetermined percent of the
CA respectively). The crossover valve duration is typically
combustion event for enhanced efficiency levels. For this
70° CA or less, and preferably 40° CA or less, in order to be
10 Study, CFD calculations showed that an overlap of 5% of theable to remain open long enough to transfer the entire mass of
total combustion event was realistic and that greater overlaps
a charge of fuel into the expansion cylinder, yet close soon
are achievable with 35% forming the unachievable upper
enough to prevent combustion from occurring within thelimit for the embodiments modeled in this study.
crossover passage. It was found that the crossover valve duraWhen the parameters are applied in the proper configuration had a significant effect on both burn rate and ITE.
IS tion, the split-cycle engine displayed significant advantages
A multiplying factor was applied to increase duration andin both brake thermal efficiency (BTE) and NOx emissions.
lift simultaneously. The valve opening point was held conWhile various embodiments are shown and describedstant, thus the valve closing event varied with duration. Since
herein, various modifications and substitutions may be madethe combustion event was held constant, an increased crossthereto without departing from the spirit and scope of theover valve duration results in a higher fraction of combustion 20 invention. Accordingly, it is to be understood that the present
occurring with the crossover valve open, which can lead toinvention has been described by way of illustration and not
flame propagation into the crossover passage for the current
limitation.
What is claimed is:
split-cycle engine configuration. Retarding the combustion
1. An engine, comprising:along with stretching the valve event would result in a sharper
25
a crankshaft rotatable about a crankshaft axis;
thermal efficiency penalty than shown here.
an expansion piston slidably received within an expansion
Stretching the valve duration and lift results in increased
cylinder and operatively connected to the crankshaft
airflow. Applying multiplying factors that result in crossoversuch that the expansion piston is operable to reciprocate
valve duration up to 42° CA, results in slight increases inthrough an expansion stroke and an exhaust stroke of aindicated power from the increased airflow. Note that thefour stroke cycle during a single rotation of the crankmultiplier for 42° CA also gives a maximum lift of3.3 mm. 30
shaft;The relationship between duration and maximum lift for FIG.
a compression piston slidably received within a compres15 is shown in table 10. For reference, the baseline configusion cylinder and operatively connected to the crankration (Run #183) had a crossover valve duration of 25° CA
shaft such that the compression piston is operable toand a maximum lift of 2.27 mm. Thermal efficiency and 35
reciprocate through an intake stroke and a compression
indicated power drop off significantly, however, with further
stroke of said four stroke cycle during a single rotation ofstretching of the valve events. Using a duration of 69° CA
the crankshaft;(and attendant increase in lift) results in 10% higher airflow,a phase angle that the expansion piston leads the compresa 9.5% drop in indicated power, and a 5.0 point drop in ITE.
sion piston by, the phase angle defined by the degrees ofTable 10 below shows the relationship between crossover 40
rotation the crankshaft must rotate after the expansion
valve duration and lift for the FIG. 29 study.
piston reaches a top dead center (TDC) position of the
expansion piston in order for the compression piston toreach a TDC position of the compression piston;TABLE 10a crossover passage interconnecting the compression cylRelationship Between Crossover Valve Duration
45
inder and the expansion cylinder, wherein the crossoverand Lift for FIG. 29 Study
passage includes a crossover valve proximate the expanCVdur
CVmax lift
sion cylinder and an inlet valve proximate the compres°CA
mmsion cylinder, the crossover valve and inlet valve defining a pressure chamber therebetween;25
2.27
Run #18350
a ratio of expansion cylinder volume from bottom dead
27.8
2.2
41.7
3.3
center (BDC) to TDC of the expansion piston is 20 to 1
55.6
4.4
or greater; and69.4
5.5
means for timing the inlet valve and crossover valve suchthat, for a period of time during which the expansion
55
piston descends towards its BDC position and the com4.0 Conclusionpression piston ascends towards its TDC position, boththe inlet valve and the crossover valve are open.
The Computerized Study identified Compression Ratio,
2. The engine of claim 1, including means for timing theinlet valve and the crossover valve such that, during a portion
Expansion Ratio, TDC phasing (i.e., the phase angle between
the compression and expansion pistons (see item 172 of FIG. 60 of the exhaust and intake strokes, both the inlet valve and the6)), crossover valve duration and combustion duration as
crossover valve are closed.
3. The engine of claim 2, wherein a substantially equalsignificant variables affecting engine performance and effimass of compressed gas is transferred from the compression
ciency of the split-cycle engine. Specifically the parameters
were set as follows:
cylinder to the crossover passage and from the crossoverthe compression and Expansion Ratios should be equal to 65 passage to the expansion cylinder as the expansion piston
or greater than 20 to 1 and were set at 100 to 1 and 120
descends towards its BDC position and the compression pisto 1 respectively for this Study;
ton ascends towards its TDC position.

US 8,006,656 B2
33

34

4. The engine of claim 2, wherein a ratio of compression

8. The engine of claim 5, including an initiator for initiating

a combustion event in the expansion cylinder while theexpansion piston is descending from its TDC position
towards its BDC position.
5. The engine of claim 4, wherein the crossover valve is an
9. The engine of claim 5, further comprising:outwardly opening valve.
a fuel injection system operable to add fuel to an exit end of
the crossover passage proximate the crossover valve.
6. The engine of claim 5, wherein a ratio of expansion
10. The engine of claim 8, further comprising:cylinder volume from bottom dead center (BDC) to top deadan exhaust valve disposed over the expansion cylinder, thecenter (TDC) of the expansion piston is 40 to 1 or greater.
exhaust valve being separate from the crossover pas7. The engine of claim 5, wherein the engine is operable 10
sage.
such that the crossover valve is open during a combustion
event in the expansion cylinder.
* * * * *cylinder volume from bottom dead center (BDC) to top deadcenter (TDC) of the compression piston is 40 to 1 or greater.