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Title:
SPLIT CYCLE SPARK IGNITION ENGINE WITH AN IMPROVED COMBUSTION CHAMBER VOLUME MODIFIER
Document Type and Number:
WIPO Patent Application WO/2018/138629
Kind Code:
A1
Abstract:
This invention relates to a split cycle reciprocating piston spark ignition engine comprising; a compressor unit to carry out the intake and compression strokes of a four stroke engine cycle, a power unit to carry out the expansion and exhaust strokes of a four stroke engine cycle, a crossover passage fluidly interconnects said compressor and power unit; the engine having substantially greater breathing efficiency by means of having greater intake aperture; a phase shift mechanism to reduce pumping loss by means of shifting the phase relation between plurality of intake valves; an expansion chamber volume modifier for enabling the engine to convert significant amount of combustion heat energy to mechanical energy during combustion.

Inventors:
MISTRY, Jiban Jyoti (F-3A, Swabhumi Apartments B G Road, Mokdumpu, Malda West Bengal 3, 732103, IN)
Application Number:
IB2018/050396
Publication Date:
August 02, 2018
Filing Date:
January 23, 2018
Export Citation:
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Assignee:
MISTRY, Jiban Jyoti (F-3A, Swabhumi Apartments B G Road, Mokdumpu, Malda West Bengal 3, 732103, IN)
SETH, Chandan Kumar (Women's College Road, PirojpurMalda, West Bengal 1, 732101, IN)
International Classes:
F02B41/00
Foreign References:
US9458741B22016-10-04
US4583501A1986-04-22
US8776740B22014-07-15
Attorney, Agent or Firm:
ANJUM IQBAL MEVEKARI (Strategic Intellectual Property Solutions, EVOMA 82/83, Borewell Road,Whitefiel, Bangalore Karnataka 6, 560066, IN)
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Claims:
CLAIMS

1. A split-cycle reciprocating piston spark ignition engine having a compression ratio of above 60: 1, the engine comprising:

at least a compressor unit (101) having a compression chamber (11) adapted to carry out the intake and compression strokes of a four stroke engine cycle;

at least a power unit (102) having an expansion chamber (21) adapted to carry out the expansion and exhaust strokes of a four stroke engine cycle;

a crossover gas passage (72) fluidly interconnecting the compression chamber (11) and the expansion chamber (21);

a first camshaft (41) and a second camshaft (42) actuating an array of fluid control valves;

a phase shift mechanism (53) phase variably interconnects the first camshaft (41) and the second camshaft (42);

an expansion chamber volume modifier (90) adapted specifically to accomplish an ignition before TDC.

2. The split-cycle reciprocating piston spark ignition engine, as claimed in claim 1, wherein said crossover gas passage (72) comprises an inlet check valve (71) at its one end, a crossover outlet valve (74) at its other end and a fuel injector (73) to inject fuel into said crossover gas passage (72).

3. The split-cycle reciprocating piston spark ignition engine, as claimed in claim 1, wherein said array of fluid control valves further comprising a pair of intake valves (70a, 70b) functional in communication with the compressor unit and at least a pair of exhaust valves (80a, 80b) and a crossover outlet valve (74) functional in communication with the compressor unit, wherein said first camshaft (41) actuates the pair of exhaust valves (80a, 80b), the crossover outlet valve (74) and a first intake valve (70a) of the pair of intake valves.

4. The split-cycle reciprocating piston spark ignition engine, as claimed in claim 3, wherein the second camshaft (42) actuates a second intake valve (70b) of the pair of intake valves.

5. The split-cycle reciprocating piston spark ignition engine, as claimed in claim 1, wherein said first camshaft (41) and second camshaft (42) are operatively connected by a phase shift mechanism (53).

6. A split-cycle reciprocating piston spark ignition engine having a compression ratio of above 60: 1, the engine comprising:

at least a compressor unit (101) having a compression chamber (11) adapted to carry out the intake and compression strokes of a four stroke engine cycle;

at least a power unit (102) having an expansion chamber (21) adapted to carry out the expansion and exhaust strokes of a four stroke engine cycle;

a crossover gas passage (72) fluidly interconnecting the compression chamber (11) and the expansion chamber (21);

a first camshaft (41) and a second camshaft (42) actuating an array of fluid control valves;

a phase shift mechanism (53) phase variably interconnects the first camshaft (41) and the second camshaft (42);

an expansion chamber volume modifier (90) adapted specifically to enable the engine to convert a significant amount of heat energy to mechanical energy during combustion.

7. The split-cycle reciprocating piston spark ignition engine, as claimed in claim 6, wherein said array of fluid control valves further comprising an array of intake valves (70a, 70b, 70c) functional in communication with the compressor unit and at least a pair of exhaust valves (80a, 80b) and a crossover outlet valve (74) functional in communication with the compressor unit, wherein said first camshaft (41) actuates the pair of exhaust valves (80a, 80b), the crossover outlet valve (74) and a pair of intake valves (70a, 70b) of the array of intake valves.

8. The split-cycle reciprocating piston spark ignition engine, as claimed in claim 6, wherein said expansion chamber volume modifier (90) comprising: a cylinder (32), a free piston (33) to reciprocate within said cylinder (32), a cylinder head (92) fitted on said cylinder (32) to form a pressure chamber (99) between said free piston (33), said cylinder (32) and the cylinder head (92); an external pressurizer (94) being fluidly connected with the pressure chamber (99) maintains a predetermined minimum fluid pressure into said pressure chamber (99).

9. The split-cycle reciprocating piston spark ignition engine, as claimed in claim 6, wherein said cylinder (32) of said expansion chamber volume modifier (90) further comprising a stopper collar (34) to secure the bottom end position of said free piston (33).

10. The split-cycle reciprocating piston spark ignition engine, as claimed in claim 6, wherein said the minimum fluid pressure maintainable in the said pressure chamber (99) is preferably set to be about 1500 kPa.

Description:
SPLIT CYCLE SPARK IGNITION ENGINE WITH AN IMPROVED COMBUSTION CHAMBER VOLUME MODIFIER

PRIORITY STATEMENT

The present application hereby claims priority to Indian patent application number 201731002954 filed 25 January, 2017, the entire contents of which are hereby incorporated herein by reference.

TECHNICAL FIELD OF THE INVENTION

The present invention relates to four stroke cycle internal combustion spark ignition engines and more specifically to a split four stroke cycle spark ignition reciprocating piston engine adapted to produce mechanical volume compression ratio of more than 60: 1.

BACKGROUND OF THE INVENTION

Ideally, higher compression ratio indicates higher thermal efficiency for a spark ignition (SI) engine. Functionally, during effective combustion in a spark ignition (SI) engine, no significant volume expansion occurs with the combustion chamber. For this reason, the combustion cycle of a SI engine is technically approximated by a "Constant- Volume Cycle". Thus, during combustion, the released heat energy may not get converted to mechanical energy (no expansion work). In a high compression ratio SI engine, this unused heat-energy may be liable for elevating the combustion chamber temperature to such a height that may create some serious problems like - large amount of NO x production, excessive heat loss to surroundings and producing abnormal combustion resulting in "knock". These are the limiting factors for existing SI engines to go for higher compression ratio and thus sacrifice the thermal efficiency as well. The compression ignition (CI) engine has comparatively better thermal efficiency as compared to the SI engines. Having approximately "constant-pressure combustion cycle" and lean burn combustion ability, the CI engine can realize higher amount of heat energy to mechanical work during combustion. These are some of the major factors that enable the CI engines to use higher compression ratio. Moreover, the combustion cycle of the CI engine consumes a considerable portion of expansion stroke and thus reduces pure expansion work. Additionally, since the volume expandability of combustion chamber relies merely on the expansion piston displacement which is governed by the crank position and speed, the heat release rate does not fit well with the volume expansion rate, hence optimum benefit is not attainable.

It is observed that a significantly higher thermal efficiency with an SI engine is achievable if the engine capable of attaining features such as high compression ratio, completion of combustion closer to TDC and conversion of combustion heat energy into mechanical work by means of significant volume expansion during combustion.

The above features altogether are not attainable with the known SI engines. Since the volume expansion rate of combustion chambers of existing engines do not agree well with heat release rate, rather it depends upon displacement rate of expansion piston which, being connected with a crankshaft, essentially depends on the crankshaft speed. Thus, though a fast combustion and high compression ratio is ideally preferable, in the existing engines, the implementation is limited due to the mismatch between the heat release rate and volume expansion rate of the combustion chamber.

Breathing efficiency is another criterion for all IC engines. Breathing efficiency or volumetric efficiency may indicate the maximum power capacity of an engine. Supercharger or turbochargers are often used to boost the breathing efficiency and thus produce greater power with a relatively smaller engine. These booster mechanisms, however, increase engine cost and bulk significantly.

Split four stroke cycle engines need fewer valves per cylinder to achieve similar breathing efficiency as existing engines. Split cycle engines generally comprise a pair of piston-cylinder assembly in which one piston-cylinder assembly is used for the intake and compression strokes (hereinafter also referred to as "the compressor unit") and another piston-cylinder assembly is used for the power and exhaust strokes (hereinafter also referred to as "the power unit"), wherein a crossover passage, being interconnected with the compressor unit and the power unit, transfers pressurized fluid or charge from the said compressor unit to said power unit where a combustion takes place consecutively.

Prior art split four stroke cycle engines, of the kind having a mechanical volume compression ratio of more than 60: 1, fundamentally start ignition significantly after TDC ( see "Split-cycle four-stroke engine", Branyon et al., US Patent No. 7588001). A significant constraint in the existing split cycle engines is, that virtually no phase overlapping being possible in such engines. This significantly affect the thermodynamic phases (power and exhaust strokes) of the power unit. At TDC the expansion chamber volume is too small (considering the mechanical volume compression ratio 60:1 or more) to induce any useable amount of charge before TDC. Thus the expansion piston has to move further after TDC until an adequate volume is formed to induce a desirable amount of charge into it. An ignition can only be initiated near about this point. Since combustion requires a finite time to mature, during that time a further piston displacement is spent. Thus, in the known split cycle engines a shorter portion of the stroke-length is available for a pure expansion work than conventional SI engines.

Moreover, another limitation of the existing split four stroke cycle engine is the shape of the combustion chamber. At the point of ignition the chamber is too shallow and wide, it may be difficult to accomplish a proper combustion due to flame quenching.

US Patent No. 9458741, titled "Split Cycle Phase Variable Reciprocating Piston Spark Ignition Engine" of Mistry discloses an engine including an expansion chamber volume modifier. The expansion chamber volume modifier comprising a free piston and cylinder assembly that makes the engine capable of starting ignition closer to TDC than the prior art split cycle engines but the engine still fires after TDC. Firing after TDC, more or less, invites some constraints like inability to run at high RPM which reduces peak power performance, produce lower peak combustion pressure that reduces work extraction ability. Moreover, a retarded ignition and combustion reduce expansion work which results in considerable loss of fuel energy to exhaust.

Further, the above patent discloses a combination of expansion chamber volume modifier and phase altering mechanism in order to achieve an early induction of fresh charge in expansion cylinder as well as to control the effective compression ratio by means of altering phase relation between said compressor and the power unit. The phase altering mechanism, however, increase the introduction difficulties like higher manufacturing cost, increase in engine mass, packaging issues, higher operating expenses and complex control algorithm to operate the system.

Another Patent. No.: US 7,588,000 B2, titled "Free Piston Pressure Spike Modulator For Any Internal Combustion Engine" of Crower et al discloses an assembly of free piston-cylinder is disclosed. Crower teaches a method and apparatus to modulate the peak combustion pressure spike and thus enable an SI engine to operate with diesel like fuel or allow higher turbo charging boost without exceeding ordinary gasoline cylinder pressures during combustion. This allows higher boost of the engine output without further engine reinforcement normally required. Thus, by keeping a lower peak combustion pressure, Crower et al also demonstrate a multi-fuel capacity of their technology as well as efficiency to prevent the pollution of the atmosphere by decreasing the formation of oxides of nitrogen (NO x ) in internal combustion engines.

However, Crower lacks demonstrating any significant benefit in thermal efficiency with their technology. Further, in all modern engines, the cylinder heads are crowded with four valves (two intake valves and two exhaust valves) per cylinder which is essential for enhancing breathing efficiency. Crower lacks demonstrating a practicable packaging solution for their device without sacrificing breathing efficiency.

Pumping loss is a significant and common constraint in SI engines. The load control in a SI engine is done by quantitative control of air-fuel mixture. This is done by obstructing the intake air flow and thus causing partial vacuum in the intake manifold. The piston needs energy to move by overcoming this vacuum. The pumping loss is greater at low load operating condition which is the most common operating condition of a vehicle. Some modern SI engines adapted Atkinson cycle (late intake valve closing) or Miller cycle (early intake valve closing) technology to alleviate the pumping loss. However, the existing mechanisms are quite complex, need precision machining and increase engine cost and bulk significantly. Thus, there is a need for simpler and more cost effective method to reduce the pumping loss.

Since, split cycle engines need fewer valves per cylinder, this creates scope for innovative and practicable application of above specified "combustion chamber volume modifier" for achieving substantially higher thermal efficiency than existing IC engines.

Accordingly, there is a need for improvement in split cycle engine technology in order to achieve significantly higher thermodynamic efficiency, lower NOx emission, while achieving reduction in manufacturing cost and design complexity. OBJECT OF THE INVENTION

An object of the invention is to provide a split cycle spark ignition engine, having a geometric volume compression ratio of about more than 60: 1, wherein the engine may be capable to produce significantly higher thermodynamic efficiency over the existing internal combustion (IC) engines by providing a four stroke internal combustion engine having at least a pair of piston-cylinder assembly, wherein the first assembly is a compressor unit which is used for the intake and compression strokes and the second assembly is a power unit which is used for the combustion cum expansion and exhaust strokes of a four stroke cycle. The compressor unit preferably intakes and compresses only air and therefore the compressed air is delivered to the power unit through a crossover gas passage. The gas passage comprises of an inlet valve on its one end connecting the compressor unit and an outlet valve on its other end connected to the power unit. Fuel is injected into the gas passage where it mixes with compressed air and the fuel-air mixture is then delivered into the combustion chamber of the power unit where subsequently a combustion is initiated by a sparkplug. An improved combustion chamber volume modifier is provided with the power unit for enabling the engine an early induction of fresh charge and particularly enabling the engine starting ignition before TDC.

An object of the invention is to provide a split cycle spark ignition engine with a novel and innovative combustion cycle, wherein a combustion of fuel-air mixture may be completed closer to TDC than conventional SI engines and at the same time the engine may attain significant expansion work during combustion.

Another object of the invention is the provision of the new combustion cycle that enables the engine for conversion of greater amount of heat energy to mechanical work during combustion and thus producing substantially higher peak combustion pressure at significantly lower combustion temperature than the existing SI engines. A further object of the invention is the provision of the combustion chamber volume modifier for enabling the combustion chamber volume to expand in response to instantaneous chamber pressure, instead of being dependent on the crankshaft speed.

A still further object of the invention is to reduce pumping loss by accomplishing Atkinson cycle (late intake valve closing) by means of shifting phase relation between a multiple intake valves.

A furthermore object of the invention is to accomplish optimum torque through a significantly greater rpm range than existing IC engines.

Another object of the invention is to provide a split cycle spark ignition engine, having geometric volume compression ratio of 80:1 or higher, capable of starting ignition before TDC along with a very quick combustion and thus the engine may be operated at higher RPM than the prior art split cycle engines.

A furthermore object of the invention is to provide a split cycle spark ignition engine that, by reducing the cycle temperature and residence time of peak cylinder temperature, may reduce or eliminate the necessity of a three way catalytic converter.

BRIEF DESCRIPTION OF THE DRAWINGS

These and other features, aspects, and advantages of the example embodiments will become better understood when the following detailed description is read with reference to the accompanying drawings in which like numerals represent like parts and sometimes like numerals followed by different alphabets represent plurality of like parts throughout the drawings, wherein:

FIG. 1 is a partially cutaway and dismantled view of the schematic of one embodiment of the engine, according to the aspects of the present invention; FIG. 2A illustrates a basic schematic plan view of the valve arrangement for the engine of FIG. 1;

FIG. 2B illustrates a basic schematic plan view of the valve actuation mechanism for the engine of FIG. 1

FIG. 3 illustrates a basic schematic of squish region provided with the cylinder head of the power unit;

FIG. 4 is a partially cutaway dismantled view of the engine illustrates the transmission scheme of working fluid from the compressor unit to the power unit of the engine of FIG. 1;

FIG. 5 is a dismantled and partially cutaway view of the engine of the present invention that illustrates the phase relation between the compressor unit and the power unit along with the combustion chamber volume at the point of ignition;

FIG. 6 is a schematic cutaway view of the engine illustrating the volume development of the combustion chamber at the point of peak combustion pressure;

FIG. 7 illustrates a schematic plan view of the valve actuation mechanism for a preferred embodiment of the present invention;

FIG. 8 is a comparative efficiency/load diagram of the engine of present invention and existing SI engine;

FIG. 9 is a comparative pressure versus crank angle (ρ-θ) diagram of the engine of present invention and existing SI engine;

FIG. 10 is a comparative temperature versus crank angle (T-θ) diagram of the engine of present invention and existing SI engine. DETAILED DESCRIPTION OF EXAMPLE EMBODIMENTS

With reference first to FIG. 1, a split cycle spark ignition engine 100 with an improved combustion chamber volume modifier 90, a first piston cylinder configuration 101 for carrying out the intake and compression strokes of a four stroke engine cycle and a second piston cylinder configuration 102 for carrying out the expansion and exhaust strokes of a four stroke engine cycle. The first piston cylinder configuration 101 may hereinafter be referred to as the Compressor Unit 101 and the second piston cylinder configuration 102 may hereinafter be referred to as the Power Unit 102.

The Compressor Unit 101 comprises a compression cylinder 20 into which a piston 12 reciprocates within a distance determined by a first crank throw 61 of a crankshaft 60 and the Power Unit 102 comprises an expansion cylinder 30 into which a piston 22 reciprocates within a distance determined by a second crank throw 62 of the crankshaft 60. A connecting rod 40 connects the piston 12 to the first crank throw 61 and a connecting rod 50 connects the piston 22 to the second crank throw 62 of the crankshaft 60. A cylinder head 31 is attached on the top of the cylinders 20 and 30. The cylinder 20 and 30, the cylinder head 31, pistons 12 and 22 typically form volume variable working chambers 11 and 21 respectively. The working chamber 11 may sometimes be referred to as the compression chamber 11 and the working chamber 21 may sometimes be referred to as the expansion chamber 21 or the combustion chamber 21. Mechanical volume compression ratio of each of the compression chamber 11 and the expansion chamber 21 is configured to be about 80: 1 or more. The cylinder head 31 comprises at least an intake port 78, an intake valve 70, a crossover inlet check valve 71 in close proximity of the cylinder 20 of the compressor unit 101 and an exhaust port (not shown), an exhaust valve 80 (shown in FIG. 2)), a crossover outlet valve 74 in close proximity of the expansion cylinder 30 of the power unit 102. The inlet check valve 71 and the outlet valve 74 are fluidly connected there-between by a crossover gas passage 72. The crossover gas passage 72 is mounted with a fuel injector 73 for injecting fuel into the gas passage 72.

Compressed air is initially delivered from compression chamber 11 to the crossover passage 72 where fuel is injected into the compressed air and thus forms combustible mixture (charge) which is then delivered to the combustion chamber 21 where consequently an ignition is initiated by a sparkplug 75. The cylinder head 31 is provided with a combustion chamber volume modifier 90 in close proximity of the cylinder 30 of the power unit 102. The combustion chamber volume modifier 90 comprises a cylinder 32, a cylinder head 92 and a free piston 33 for reciprocating within the cylinder 32. The free piston 33 has two working ends, i.e. a bottom end 33a and a top end 33b. A stopper collar 34 rigidly mounted with the cylinder 32 to secure the bottom end position of the free piston 33. The top end 33b of the free piston 33, the cylinder 32 and the cylinder head 92 forms a pressure chamber 99. The bottom end 33a of the free piston 33 is exposed to the combustion chamber 21 for modifying expansion chamber volume and shape. The displacement of free piston 33 is governed by an instantaneous pressure difference of combustion chamber 21 and the pressure chamber 99.

The combustion chamber volume 21 thus includes a first combustion chamber volume 21a formed by displacement of free piston 33 and a second combustion chamber volume 21b formed by displacement of the expansion piston 22. The cylinder head 92 is provided with an inlet check valve 94 to allow one way flow of pressurized air from an external source to the pressure chamber 99 in order to maintain a predetermined minimum pressure in said pressure chamber 99. The said minimum pressure is determined by the pressure of the pressure chamber 99 when it is in fully expanded condition, i.e., the piston 33 is at its lowest position or the bottom end 33a of piston 33 is aligned with a reference line 35 (shown by phantom line 35). Wherein, said predetermined minimum pressure value of pressure chamber 99 is less than an induction pressure of the combustion chamber 21. A pressure of about 15 bar is set to be a preferred minimum pressure maintainable in the pressure chamber 99.

Further, FIG. 1 shows an intake and an expansion phase in the compressor 101 and the power unit 102 respectively. The compression piston 12 descends through an intake stroke and the expansion piston 22 descends through an expansion stroke. The downward motion of compression piston 12 creates partial vacuum in chamber 11, when in synchronization, the intake valve 70 opens by an actuator 79 to allow induction of fresh air into said chamber 11. The expansion piston 22 is descending through an expansion stroke which causes pressure drop in the expansion chamber 21. The free piston 33 also descends due to pressure differential between expansion chamber 21 and pressure chamber 99. Thus, the aggregated volume of expansion chamber 21 and pressure chamber 99, which is substantially larger than the expansion chamber alone, contribute to the expansion work until the free piston 33 reaches to its lowest position. This phenomenon substantially enhances the thermodynamic efficiency over prior art engines.

FIG. 2A illustrates a basic schematic of cylinder head 31 of the split cycle engine of the present invention. Cylinder head 31 communicates with the compressor unit 101 and the power unit 102. A first portion 31a of cylinder head 31 communicates with the compressor unit 101. Said first portion 31a generally comprises a first intake valve 70a, a second intake valve 70b and a crossover inlet check valve 71. A second portion 31b of cylinder head 31 communicates with the power unit 102. Said second portion 31b generally comprises a cylinder 32 in which reciprocates a free piston 33, at least a pair of exhaust valves 80a, 80b and a crossover outlet valve 74 and plurality of sparkplugs 75a and 75b. The crossover inlet check valve 71 and the crossover outlet valve 74 are fluidly connected by a crossover passage 72 (shown in phantom lined arrow 72).

FIG. 2B illustrates a control scheme of the valves as illustrated in FIG. 2A, wherein a first camshaft 41 comprises plurality of cam lobes 43, 46, 47 and 48 of different cam profile. Said first camshaft 41 is configured to actuate the first intake valve 70a, the pair of exhaust valves 80a, 80b and the crossover outlet 74. A second camshaft 42 includes a cam lobe 45 having a cam profile similar to cam lobe 43 of said first camshaft. Said second cam lobe 45 being mounted with the second camshaft 42 actuates the second intake valve 70b. The first and second camshafts are operatively connected between them by a phase shift mechanism 53. At low- load operating condition the phase shift mechanism 53 is configured to retard the phase angle of the second camshaft 42. This produces late opening and closing of the second intake valve 70b. Thus the intake valve 70b is kept open during a portion of compression stroke and thereby a fraction intake fluid may get back to a corresponding intake manifold. The amount of phase shift may be determined by considering instantaneous load demand. At full load, both of the intake valves 70a and 70b operate at same phase angle and thus an optimum induction of air may be accomplished.

FIG. 3 shows the second portion 31b of cylinder head 31, that communicates with the power unit 102 wherein, the crosshatched area shows a squish region 36. Generally, when a piston reaches its TDC, at squish region the gap between the cylinder head and the piston crown nearly disappears. During a compression stroke, as a piston approaches near TDC, the fluid within the squish region and piston crown rushes out and thus creates turbulence in the thick area of combustion chamber where a combustion occurs. Larger squish region thus generates substantially greater turbulence. Turbulence is an important factor that increase the combustion speed. The squish region 36 of the engine of the present invention is substantially larger than existing SI engines. This enables the split cycle engine of the present invention to attain a very short combustion duration.

FIG. 4 illustrates a transmission scheme of working fluid from the compressor unit 101 to the power unit 102 at about 15 crank angle degrees (CAD) before ignition in the engine 100 of FIG. 1. In an embodiment, the compression piston 12 ascends through a compression stroke and the expansion piston 22 ascends through an exhaust stroke. In this particular embodiment the compression piston 12 is configured for 5 CAD phase advanced with the expansion piston 22 i.e., the compression piston 12 reaches its top dead center (TDC) at about 5 CAD before the expansion piston 22 reaches its TDC. Pistons 12 and 22 are approached within 15 CAD and 20 CAD respectively from TDC and still ascending.

The compressor unit 101 intakes and compresses only air. The crossover passage 72 retains pressurized air from previous cycle. As the pressure of the compression chamber 11 exceeds the pressure of the crossover gas passage 72, compressed air starts to transfer from the compression chamber 11 to the crossover gas passage 72 by means of opening of the crossover inlet check valve 71. The inlet check valve 71 permits one way flow of working fluid from compression chamber 11 to crossover gas passage 72, wherein opening and closing of check valve 71 is governed by instantaneous pressure differential between the compression chamber 11 and the crossover gas passage 72. In synchronization with the transmission of compressed air into the crossover gas passage 72, the fuel injector 73 injects fuel into said passage 72, whereupon being mixed with fuel, the combustible fluid is transferred to the combustion chamber 21 by timely opening of the crossover outlet valve 74. Opening and closing of the crossover outlet valve 74 is controlled by an actuator (not shown). The exhaust valve 80 (see FIG. 2A) is configured to move at its close position before the air-fuel mixture starts transferring in the combustion chamber 21.

As a considerable amount of compressed fluid is induced to the combustion chamber 21 at a fluid pressure higher than the pressure (about 15 bar) in the pressure chamber 99, the pressure differential cause a displacement of the free piston 33 towards the pressure chamber 99 and thus a volume 21a is formed.

Referring to FIG. 5, wherein the compressor piston 12 has reached at its TDC position and the expansion piston 22 are approached within a several CAD (about 5 CAD before TDC) of its TDC and still ascending towards said TDC, whereby a major portion of combustible fluid (fuel-air mixture) is transferred to the expansion chamber 21 and a combustion is initiated about this point by a sparkplug 75. It would be apparent from FIG. 5, that at this position, the volume portion 21b of the combustion chamber volume 21 is too small that an adequate amount of combustible fluid cannot be transferred to the portion 21b, whereupon the volume would be even smaller at TDC. This generally prevents prior art split cycle engines from induction of fresh charge before TDC. In this embodiment of the present invention, the improved combustion chamber volume modifier 90 is adapted to produce an additional volume 21a in order to achieve the induction of charge before TDC and thus enabling a start of combustion before TDC.

As the induction pressure in volume portion 21b of combustion chamber 21 exceeds the pressure within the pressure chamber 99, the free piston 33 starts displacing towards said pressure chamber 99 and continues until an equilibrium pressure state between said combustion chamber 21 and said pressure chamber 99 is accomplished. Thus, at TDC the volume portion 21a becomes the major part of combustion chamber 21. Since the combustion starts before TDC, the progressive combustion pressure continues to cause further expansion of volume 21a until peak combustion pressure is accomplished. The point of ignition is preferably configured to be 5 to 10 CAD before TDC and thus a peak combustion pressure is attainable within 10 to 15 crank angle degrees after TDC.

Air-fuel mixture is transferred to the combustion chamber 21 at nearly fully compressed condition and so the mixture enters into the combustion chamber at a very high velocity and thus produces a very strong turbulence in the combustion chamber 21. This turbulence is continued thereon with a squish generated turbulence. Since, at the point of ignition the air-fuel mixture is nearly fully compressed (contrary to the conventional SI engine where ignition initiate well before reaching full compression, i.e., about 20 to 25 CAD before TDC), during ignition the mixture density, pressure and turbulence is significantly higher than the prior art SI engines. Further, because no combustion occurs in the compressor unit, the compression chamber 11 contains no residual burned gas and only air is used for compression, the compression occurs at significantly higher specific heat ratio than conventional internal combustion (IC) engines. This produces considerably higher thermodynamic efficiency than the existing IC engines.

With reference to FIG. 6, wherein the compression piston 12 has moved 17 CAD past TDC and the expansion piston 22 has moved 12 CAD past TDC. Peak combustion pressure is accomplished about this position. From the point of ignition (between 5 to 10 CAD before TDC), a nearly negligible expansion occurs in the volume portion 21b of combustion chamber 21. Whereas, in response to the elevated combustion pressure the volume portion 21a of combustion chamber 21 expands substantially. Thus, a considerable amount of heat energy that is released during combustion gets converted into mechanical energy and is stored as potential energy in the pressure chamber 99. Thus, unlike conventional engines, the combustion chamber volume expansion rate of the split cycle engine of present invention is not limited by the displacement of the expansion piston. Therefore, during combustion a significant volume expansion occurs with the combustion chamber and thus, a considerable amount of heat energy get converted into expansion work. This conversion of heat energy greatly reduces the peak combustion temperature and residence time of the peak temperature as well. Thus, necessity of a three-way catalytic converter may be reduced or eliminated.

As the expansion piston 22 moves further towards its bottom dead center (BDC) the cylinder pressure start dropping and in response of this pressure drop the pressure chamber 99 start expanding by displacing the free piston 33 towards its lowest position 35. Thus, this stored energy is recovered and distributed through a significant portion of expansion stroke of the expansion piston 22 until the free piston 33 reaches to its lowest position. FIG. 7 illustrates a schematic of a preferred alternative of the present invention, wherein the portion 31a of the cylinder head 31 further comprises an array of three intake valves 70a, 70b, 70c and a crossover inlet check valve 71 connecting one end of the crossover passage 72. The portion 31b of the cylinder head 31 further comprises a pair of exhaust valves 80a, 80b, the cylinder 32 of expansion chamber volume modifier 90, sparkplugs 75a, 75b and a crossover outlet valve 74 connecting the other end of the crossover passage 72.

A first camshaft 41 having a first pair of cam lobes 43, 44; a second pair of cam lobes 46, 47 and a fifth cam lobe 48. The first pair of cam lobes 43, 44 are provided for simultaneous actuation of two intake valves 70a, 70b of the array of the intake valves 70a, 70b and 70c. The second pair of cam lobes 46, 47 are provided for timely actuation of the pair of exhaust valves 80a and 80b respectively. The fifth cam lobe 48 is provided for actuating the crossover outlet valve 74. Said first camshaft 41 is driven conventionally through a pulley or sprocket 51 connected by a timing belt or chain 52 connected to a corresponding engine's output shaft. A second camshaft 42 having a cam lobe 45 is provided for actuating the third intake valve 70c of the array of intake valves 70a, 70b and 70c. The second camshaft 42 is operative by the first camshaft 41 as being connected by a phase shift mechanism 53.

Load control in an SI engine is essentially done by controlling the fuel- air mixture quantitatively. This has generally been done by simply throttling the intake passage. At low loads, this method cause severe pumping loss which ruins the thermal efficiency of the engine.

At low loads, the phase shift mechanism 53 retards the phase angle of the second camshaft 42 by an angle determined by considering instantaneous load demand. Thus, to induce a fraction of air-fuel mixture the third intake valve 70c is kept open during a portion of compression stroke so that a fraction of intake air may be send back to a corresponding intake manifold. The late intake valve closing method is generally known as the "Atkinson cycle" which effectively reduces the pumping loss as well as produces over expansion cycle.

In existing engines, the Atkinson cycle is generally attained by varying the opening time duration and stroke length of the intake valves. These methods are quite complex and expensive to implement. In contrast, the split cycle engine of the present invention provides a simple method to attain the "Atkinson cycle" by means of varying only the phase relation between multiple intake valves without compromising the engine's breathing efficiency.

It is evident that greater intake aperture provides higher breathing efficiency. Contrary to the two intake valves system of existing engines, the three intake valves of the engine of present invention provides 50% more intake area. This greatly widen the engine's rpm band for achieving high torque.

Several mathematical analysis have been carried out in order to predict the comparative thermal efficiency between the split cycle SI engine of the present invention and the existing SI engines at different engine loads. It is revealed from the analysis that the capability of significant volume expansion of the combustion chamber during combustion enables the split cycle engine to attain significantly higher compression ratio than existing SI engines. This is because a significant amount of combustion heat energy get converted to mechanical energy due to the expansion work and thus keeps the combustion chamber temperature significantly lower than existing SI engines. It is also observed that, because the volume expansion rate of the combustion chamber of the split cycle engine is not limited by the crankshaft speed, rather it rely on instantaneous combustion pressure, high torque can be attained at a very low rpm also. Since the present split cycle engine has 50% more intake aperture it can breathe well at substantially higher rpm and thus can also attain high torque at substantial higher rpm than existing engines. The above features greatly widen the torque-rpm band of the split cycle engine of the present invention. FIG. 8 shows a comparative efficiency/load diagram of the present split cycle engine and existing SI engine. The efficiency trace of split cycle engine represents a substantially greater thermal efficiency than existing engines throughout the load conditions. It can be seen from the diagram that, unlike existing SI engines the thermal efficiency of the split cycle engine do not decrease with decreasing engine load, rather at the most common drive conditions the thermal efficiency is considerably higher than full load efficiency.

FIG. 9 shows a comparative cylinder pressure trace during expansion stroke. The pressure-crank angle (ρ-θ) diagram of the split cycle engine and existing SI engine at 0.3 load is compared. The pressure trace of split cycle engine represents a substantially greater thermal efficiency than existing SI engines.

FIG. 10 shows a comparative temperature-crank angle (T-θ) diagram of present split cycle engine and existing SI engine during expansion stroke at the most common part load (0.3 load) operating condition. The temperature trace of split cycle engine shows a substantially lower cylinder temperature than existing engine. The pressure and temperature traces of FIG. 9 and 10 together reciprocally represent the substantially greater energy conversion efficiency of the split cycle engine of the present invention over the prior art SI engines.

Moreover, the engine provides a new and highly efficient combustion cycle wherein both of the features like 'combustion closer to TDC and 'significant expansion work during combustion' are attainable simultaneously. This is the key factor for achieving a substantially higher thermal efficiency than existing SI engines.

As will be understood by those skilled in the applicable arts, various modifications and changes can be made in the invention and its particular form and construction without departing from the spirit and scope thereof. The embodiments disclosed herein are merely exemplary of the various modifications that the invention can take and the preferred practice thereof. It is not, however, desired to confine the invention to the exact construction and features shown and described herein, but it is desired to include all such as are properly within the scope and spirit of the invention disclosed.