EP0476010B1

EP0476010B1 - Reciprocating piston engine with pumping and power cylinders

Info

Publication number: EP0476010B1
Application number: EP90909109A
Authority: EP
Inventors: Glen Allan Dullaway
Original assignee: ROTEC ENGINES Pty Ltd
Current assignee: ROTEC ENGINES PTY. LTD.
Priority date: 1989-06-16
Filing date: 1990-06-15
Publication date: 1995-08-23
Anticipated expiration: 2010-06-15
Also published as: EP0476010A1; AU5818490A; WO1990015917A1; EP0476010A4; AU638720B2

Abstract

An internal combustion engine power unit comprises two power cylinders (3, 4) spaced equidistant about a pumping cylinder (5). All cylinders operate on two-stroke cycles, the power cylinders (3, 4) having a phase difference of 180°. Power piston asssemblies (13, 14) in the power cylinders (3, 4) drive crankshaft (1). Pumping piston (16) and separate crankshaft (2) are driven at twice the cyclic speed of the power pistons (13, 14) and crankshaft (1) through gear train (6, 7) between the respective crankshafts (1, 2). Air inducted into pumping cylinder (5) via intake ports (20) is compressed and passed alternately to power cylinders (3, 4) via valve controlled transfer passages (21, 24). All valves, ports and gas passages are found in a cylinder head (19). Timed fuel injection and ignition are provided. An engine may comprise one or more power units. There is also disclosed a turbo-charged diesel engine comprising two power units in ''V'' configuration.

Description

TECHNICAL FIELD

This invention relates to reciprocating piston internal combustion engines of the type wherein, pumping and power cylinders are operated on two stroke cycles.

BACKGROUND ART

Engines of this type have been disclosed in numerous prior art which have intended to improve the efficiency and or power to weight ratio thereof. US,A, PATENT NO 1,881,582 shows a design which has a pumping cylinder driven at twice the cyclic speed of and alternately suppling a intake scavenging charge to two power cylinders, via transfer ports which communicate with the lower cylinder walls of the power cylinders, hence being timed by the power pistons. Although this design marginally increases the scavenging efficiency attainable and as compared to crankcase compression type two stroke engines, this design has and retains numerous efficiency problems of, including the fundamental inefficiency of, the conventional two stroke engine. The said inefficiency results from the opening of the transfer ports in the lower cylinder walls and which reduces the volume through which expansion occurs with the said reduction being used instead for a half of the transfer scavenging phase. Furthermore this design, due to the said transfer to the lower cylinder walls, has no potential for significant efficiency gains to be attained if valve controlled constant volume combustion chambers are to be used.
A second type of engine which has pumping and power cylinders operating on two stroke cycles and which have intended to overcome the above said undesireable features are typically disclosed in U S PAT NO'S : 3,880,126 and 4,458,635. These designs have the pumping cylinder transferring the intake charge through valve timed ports which open into the power cylinder head section. U S PAT NO : 3,880,126, utilizes a combustion chamber which is in constant communication with the power cylinder and which has an excessive number of components whilst overall efficiency and power output are severley limited by a poor scavenging efficiency which primarily results from the long transfer scavenging phase required of the design. This further exacerbats the obvious power to weight ratio limitations of the design. U S PAT NO: 4,458,635, utilizes a valve controlled constant volume combustion chamber which foregoing the supercharging system used that results in a similar said fundamental inefficiency, increases the scavenging and combustion efficiency and hence overall efficiency is also maginally increased. Subsequently, only an average power to weight ratio results whilst an excessive number of components is still a major problem.
A further design of engine which shares similiar cylinder, port and valve locations of the presented invention but which is outside of the technical field of this invention in that the power cylinders operate on four stroke cycles, is typically shown in GB,A PATENT NO 2071210. As such the pumping cylinder is used only as a supercharging device and is not necessary for the operation of the engine as is required in the presented invention.
As acknowledged in the pre-characterising portion of claim 1 appended hereto, French Patent No. 820925 to GRAB discloses a two-stroke internal combustion engine comprising at least one unit having a pumping cylinder, a pumping piston reciprocally movable in said pumping cylinder, two power cylinders, a respective power piston reciprocally movable in each said power cylinder, each said power cylinder having an associated combustion chamber, the pumping piston reciprocating at a cycle speed twice that of the power pistons, and said power pistons being phased about one stroke apart, a cylinder head closing the top ends of all said cylinders, said cylinder head having two transfer ports therethrough enabling said pumping cylinder to communicate with said power cylinders, transfer valves controlling communication between the pumping cylinder and the power cylinders, at least two exhaust ports allowing exhaust gases to flow from the power cylinders when the associated power piston is at about or before its bottom dead centre position, at least one intake port through the head and communicating with the pumping cylinder, intake valve means associated with the intake port and allowing a major portion of an intake charge to be induced into the pumping cylinder when the pumping piston is moving away from its top dead centre position and said pumping piston alternately transferring the intake charge into the power cylinders through the transfer ports as the pumping piston moves towards its top dead centre position, said pumping piston leading to the top dead centre position the power piston of the cylinder to which the charge is transferred.
The fixed geometry of the ported portion of the GRAB engine means that exhaust timing is dependent upon piston position such that opening and closing events are interdependent and exhaust open timing cannot be decided without consideration to duration or closing. Further, the exhaust timing is necessarily restricted to being symmetrical about BDC; as such, exhaust timing is compromised. Also, the engine of GRAB employs an axial flow scavenging motion.
A further problem of the GRAB engine is that its exhaust timing cannot be controlled and manipulated in operation as the engine is traversing its load range. Thus, for example, the scavenging and trapping efficiencies of the engine, that is, the amounts of fresh charge and retained exhaust gas trapped in the cylinder ready for combustion, cannot be controlled and manipulated by control of the exhaust opening and closing events.
More generally, the ports used by GRAB require the cutting of holes in the cylinder walls. These holes would be detrimental to the life of the piston ring and would cause premature wear and degradation of the sealing efficiency of the engine leading to early failure and increased pollution throughout the engine's life. This point is critical in view of increasingly stringent exhaust emissions legislation around the world. For example, the US Clean Air Act 1990 requires that vehicle engines must endure at least 100,000 miles (160,000 kilometres) from new without significant increase in tailpipe emissions. It is believed that, to date, no ported engine has succeeded in passing these requirements.
The present invention is characterised in that: the flow of the exhaust gases through the exhaust ports is controlled by exhaust poppet valves; the exhaust ports extend through said head; the exhaust valves begin to open when the power piston is positioned between 90 degrees after top dead centre and 198 degrees after top dead centre and close when the power piston is positioned between 162 degrees before top dead centre and 10.8 degrees before top dead centre; the transfer valves are poppet valves; the transfer valves begin to open when the pumping piston is positioned between 72 degrees after top dead centre and 288 degrees after top dead centre and close when the pumping piston is positioned between 72 degrees before top dead centre and 72 degrees after top dead centre; each exhaust valve closes before the associated transfer valve closes; the intake valve means begin to open when the pumping piston is positioned between top dead centre and 144 degrees after top dead centre and close when the pumping piston is positioned between 216 degrees before top dead centre and 14.4 degrees before top dead centre; and the pumping piston leads the power piston to which the intake charge is to be transferred to the top dead centre position by less than 108 power piston degrees before top dead centre.
The invention provides a novel engine design which significantly increases the thermal efficiency, power to weight ratio and scavenging efficiency of all the abovesaid engine types. Further, the number of components required for the abovesaid second type of engine is reduced.
Advantageous optional features of the invention are set forth in claims 2 to 17 appended hereto.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG 1- is a top schematic view of the preferred design which is a inline single unit and showing the cylinders, ports, combustion chamber and valve opening locations thereof.
FIG 2- is a cross-sectional view taken along line A-A of FIG 1 but around the piston crankshaft mechanism and with the lower crankcase removed.
FIG 3- is a valve timing diagram of the preferred design in power cylinder crank angle degrees with the lines indicating valve open times and with the TDC position shown thereon being the TDC position of the first power piston.
FIG 4- shows an alternative design which has two units, being in a V configuration and utilizing turbocharging and crankcase compression of the pumping cylinder. One unit or bank of cylinders is shown as an end view with the other unit shown as a cross-sectional view taken along line B-B of FIG 5 but around the piston crankshaft mechanism thereof and with partial hidden detail shown and the lower power crankcase and lower RH side pumping cylinder crankcase being removed.
FIG 5- is a top schematic view of the sectioned unit of FIG 4 and shows the cylinders, ports, combustion chambers and valve opening locations thereof.
FIG 6- is a valve timing diagram of the alternative design shown in FIGS 4 and 5 and uses the same features as described for FIG 3.
FIG 7- is a end shematic view of an alternative V configuration and which shows the cylinder and crankshaft locations thereof.

MODES FOR CARYING OUT THE INVENTION

Refering to all modes for carrying out the present invention, each said unit has a pumping cylinder 5 with a pumping piston 16 reciprocable therein and first and second power cylinders, respectively 3 and 4 with first and second power pistons respectively 13 and 14 reciprocable within their respective power cylinders. All cylinders of a unit share a parallel axis and a common block 18 and a common head 19, whilst the pumping cylinder is evenly distanced to each of the power cylinders. A pumping crankshaft 2 and pumping conrod 17 cause reciprocation of the pumping piston 16 and a power crankshaft 1 and power conrods 15 cause reciprocation of the said power pistons. Each said crankshaft is supported for rotation by bearing means whilst journal means which are not shown in the drawings, provide pivotal movement at the conrod crankshaft pivots and the conrod piston pivot. A pump drive gear 7 which is fixed to each of the pumping crankshafts 2, cooperates with, is driven by, and is one half the diameter of, the power crankshaft gear 6 which is fixed to the power crankshaft 1. This gear arrangement then provides for the pumping pistons 16 to be reciprocated at and cyclicly operated at, twice that of the power pistons. The phasing of the power pistons of a unit relative to each other, is one hundred and eighty power 'crankshaft crankangle' (from hereinafter is refered to as 'CA') degrees. The first and second transfer ports respectively 21 and 24, remain in constant communication with the pumping cylinder. The crankshafts for carrying out all modes of the invention, are of the one piece type whilst all conrods are of the two piece type and bolt on to the respective crankshafts from the underside thereof, for pivotal movement therearound. Of course the components and auxiliaries not illustrated for the efficient operation of the engine, are included in all modes for carrying out the invention whilst water cooling passages are shown in the sectioned walls of FIGS 2 and 4 but are not numbered to reduce cluttering thereof. Furthermore, the respective components of the first and second power cylinders are respectively referred to as the first and second said components, or they are referred to as the respective components of the power cylinder of which the description is directed to.
Referring now to FIGS 1-3, the preferred design or mode of carrying out the invention is a naturally aspirated inline version with the pumping cylinder 5 being located in the middle of the first and second power cylinders, respectively 3 and 4. The pumping crankshaft 2 is accessed and held in place by pumping crankshaft caps 38 which bolt into the engine block 18 whilst the power crankshaft 1 is accessed and held in place by the lower crankcase which is removed in the FIG 2. The phasing of the pumping piston 16 relative to the power pistons 13 and 14 has the pumping piston leading the piston of the power cylinder which the intake charge of that particular pumping cylinder cycle will be transferred into, to TDC, by less than 108 power CA degrees.
The preferred design has all intake, transfer, and exhaust valves being poppet type valves. The first and second combustion chambers respectively 22 and 25, remain in constant communication with their respective power cylinder and each has a spark plug 35 mounted thereinto and which causes ignition of the combustible mixture therein. Petrol fuel injection means 36 are mounted into each transfer port and inject a predetermined quantity of fuel thereinto as the said intake charge is being transferred into the power cylinder thereof. A first transfer valve 8 times communication between the first transfer port 21 and the first power cylinder 3 whilst a second transfer valve 10 times communication between the second transfer port 24 and the second power cylinder 4.
Two intake valves 12 time communication between the intake port 20 and the pumping cylinder 5. A first exhaust valve 9 times communication between the first power cylinder 3 and the first exhaust port 23 whilst a second exhaust valve 11 times communication between the second power cylinder 4 and the second exhaust port 26. The said exhaust ports lead to an exhaust manifold and eventually to an exhaust pipe whilst the said intake port leads to an intake manifold with air metering means therein provided. All of the said valves are actuated by a single overhead camshaft which has an axis parallel to that of the crankshafts and is positioned directly above all the said valves so as to directly actuate them. The said camshaft is not shown in FIG 2 to reduce cluttering thereof and of the major features thereof. The said camshaft is driven by chain means 46 from the camshaft drive sprocket 39 which is fixed to the pumping crankshaft 2. The sprocket on the said camshaft which cooperates with the said chain is a half of the diameter as the said camshaft drive gear, providing for the said camshaft to cooperate at the same cyclic speed as the power cylinders and as such, single camlobes actuate the transfer and exhaust valves, whilst two camlobes are evenly spaced around the said camshaft where the intake valves are actuated from, so that the intake valves open twice as often as the other valves and which follows the increased cyclic speed of the pumping cylinder. Variable exhaust valve closing event is obtained by a turning block type of valve timing mechanism which is not shown for reasons of undue complexity and which allows for the said valves to close between 162 and 10.8 power CA degrees before TDC depending upon engine load and speed. The engine oil pump supplies the oil to the engine and is driven from the oil pump drive gear 40 which is fixed to the power crankshaft 1 between the power cylinders.
The method of operation including the valve timings of the preferred design is now described. The intake valves 12 open when the pumping piston moves between 0 and 144 pumping CA degrees after TDC. This allows the compressed intake gas of the previous cycle to expand substantially to atmospheric before the said valves 12 are opened. With the intake valves 12 opened and the pumping piston moving towards its 'bottom dead centre' (which hereinafter is referred to as 'BDC') position, the intake air is induced into the pumping cylinder 5. As the said piston 16 moves through 36 pump degrees before BDC the intake valves 12 close and the induction of the intake air ceases. One of the transfer valves 21 or 24 begins to open between 72 and 288 pump degrees after TDC, initiating the transfer phase to the respective power cylinder of which the said open transfer valve opens into. The said transfer valve then remains open until the pumping piston 16 moves through to 72 pump degrees after TDC which is shown in FIG 3. The piston of the pumping cylinder then continues towards BDC, and begins a new cycle thereof as is described above. The intake air of the next said cycle is transferred to the other power cylinder and the intake air of the following said cycle and which is after the next said cycle is transferred to the said respective power cylinder starting a new cycle thereof.
During the first part of the transfer phase to the said respective power cylinder, the exhaust valve thereof is open, providing for the later part of the exhaust phase thereof to occur which has the scavenging of the remaining exhaust gases from the said respective cylinder by the transferring intake air. The exhaust valve of the said respective power cylinder remains open until the piston thereof moves to between 162 and 10.8 power CA degrees before TDC. At high load and or high speed, the fuel is injected into the transfer port of the said respective power cylinder during the transfer phase and at low load and or speed, it is mostly injected after the exhaust valve of that power cylinder has closed. With the fuel injected, a spark at the respective spark plug 35 causes combustion to occur about the TDC position. The piston of the said respective power cylinder then moves towards BDC, substantially expanding the gases therein to atmospheric before the exhaust valve begins to open when the said piston is between 90 and 198 power CA degrees after TDC. This then initiates the first part of the exhaust phase being blowdown, and then positive scavenging occurs whilst the piston thereof moves towards TDC until the transfer valve of that cylinder opens, beginning another cycle thereof and as is described above. The operation of the other power cylinder is the same as that described above for the said respective power cylinder but as is obvious, it occurs one hundred and eighty power CA degrees before and after it occurs in the said respective cylinder.
Referring now to FIGS 4-6, the alternative design or mode for carrying out the invention has two units which are set in a V configuration and with each said unit being one bank of cylinders of the said V. The power cylinders of each unit, are positioned close together with the pumping cylinder 5 of each unit being positioned on the outside of the said V but being central to the power cylinders of its said unit. Constant volume combustion chambers which have communication to their respective power cylinders being timed by secondary valves are used in the alternative design with the first said secondary valve being 27 and the second said secondary valve being 28. A turbocharger 41 is positioned in the middle of the said V with the exhaust manifolds 23 of all power cylinders communicating thereto whilst the exhaust, ports 23 and manifolds 23 share the same number. The pressurised intake manifold 42 leading from the turbocharger 41 communicates with the intake ports of both pumping cylinders whilst the crankcase intake ports 33 of both pumping cylinders is naturally aspirated. A single power crankshaft 1 causes reciprocation of all power pistons whilst each pumping cylinder 5 has its own pumping crankshaft 2. A single power crankshaft gear 6 which is fixed to the power crankshaft, cooperates with the pumping cylinder drive gears 7, fixed to each of the pumping crankshafts. The phasing of the pumping pistons relative to the power cylinders of a respective unit, has the pumping piston leading the said power pistons to TDC by less than 108 power CA degrees. The phasing of the power pistons of the unsectioned unit relative to the said pistons of the sectioned unit, has the first power piston 3 of the sectioned unit, leading the first power piston of the unsectioned unit, by ninety power CA degrees. The power crankshaft 1 is accessed and held in place by the lower power crankcase which is removed in the drawings whilst each pumping crankshaft 2 is accessed and held in place by a pumping lower crankcase 47 which is shown on the unsectioned unit of FIG 4.
The alternative design has all intake, transfer, exhaust and secondary valves being poppet type valves whilst the crankcase intake valves 32 are reed type valves. The first combustion chamber 22 and the first power cylinder 3 has communication therebetween controlled by a first secondary valve 27 whilst the second combustion chamber 25 and the second power cylinder 4 have communication therebetween comtrolled by a second secondary valve 28. Diesel fuel injection means 37 are mounted into each said combustion chamber whilst ignition therin is caused by the temperature and pressure of the combustible mixture therein. Protrusions 31, on the top of each power piston, extend upwards so that they substantially atleast, take up the volumes of each secondary port 29 and 30 which result in an efficiency increase of the engine. The alternative design has each unit having the same intake, transfer, and exhaust valve and port arrangements and functions, as are described for the preferred design although the positioning of some valves and ports is altered. Each said unit has two overhead camshafts which are not shown in the drawings and which are driven by gear means from the pumping cylinder drive gear 7. One of two idler gears 43 cooperates with the said gear 7 whilst the another idler gear 44 cooperates with the idler gear 43 and with the camshaft gear 45 which is the same diameter as the power crankshaft gear 6. The said camshaft gear 45 is fixed to the power camshaft which has single camlobes actuating each transfer, secondary, and exhaust valves whilst another gear which is fixed to the said power camshaft cooperates with a gear which is a half the diameter thereof and which is fixed to the pumping camshaft. The said pumping camshaft has single camlobes actuating the intake valves with the said diameter difference of the relevant gears providing for the increased cyclic velocity of the intake valves.
The method of operation including the valve timings of the alternative mode is now described with reference to a single unit. The intake valves 12 begin to open when the pumping piston moves between TDC and 144 pumping CA degrees after TDC. This allows the compressed air from the previous cycle to expand substantially to the pressure of the intake manifold before it opens. With the intake valves 12 opened and the pumping piston moving towards its BDC position, the intake air is induced into the pumping cylinder 5. Whilst the said piston 16 is moving towards BDC, the intake air within the crankcase is compressed. If the engine is operating above or about, fifty percent of its possible load, then the turbocharger 41 is operating efficiently, and as the pumping piston uncovers the crankcase transfer ports 34 between 72 and 288 pump degrees after TDC, then no crankcase transfer occurs as the pressure in the said cylinder resulting from the turbocharger 41 is as high or higher than that of the said crankcase. This then provides for the said crankcase compression to be utilised at the lower loads but not at the higher loads as well as minimising the maximum pressures attained in the said crankcase which then reduces the sealing requirement thereof and allowing for lighter said reed valve materials with lower opening pressures. As the said piston 16 moves between 72 pump degrees before TDC and 72 pump degrees after TDC, the crankcase transfer ports 34 are closed and when the said piston moves between 216 and 14.4 pump degrees before TDC, the intake valves 12 are closed and the induction of the intake air through the intake ports ceases whilst if the engine is operating at low load then on the said pistons up stroke, intake air will be induced into the crankcase through the crankcase intake valves 32. One of the transfer valves opens when the pumping piston is between 72 and 288 pump degrees after TDC, to initiate the transfer phase to the respective power cylinder which the said transfer valve opens into. The said transfer valve then remains open until the pumping piston has moved between 72 pump degrees before TDC and 72 pump degrees after TDC. The piston of the pumping cylinder then continues towards BDC and begins a new cycle thereof when the intake valves begin to open again between TDC and 144 pump degrees after TDC whilst the intake air of the next said cycle is transferred to the other power cylinder and so forth as is described hereinbefore.
During the first part of the transfer phase to the respective power cylinder, the secondary valve thereof is open providing for the scavenging of the exhaust gases from the combustion chamber thereof. The said secondary valve closes when the piston of that respective power cylinder has moved to one hundred and fifteen power CA degrees before TDC. During this time the exhaust valve of the said respective power cylinder is open and closes when the piston thereof has moved to forty five power CA degrees before TDC, allowing for nearly all the exhaust gas to be scavenged from the said cylinder except for a small residual portion thereof remaining. This is retained to highly pressurise the remaining gas so that when the secondary valve reopens when the piston thereof is at five power CA degrees before TDC, the pressure in the power cylinder is not significantly lower than that of the combustion chamber thereof which would decrease the thermal efficiency attainable. When the said piston is positioned about forty power CA degrees before TDC, diesel type fuel is injected into the said combustion chamber which results in combustion occuring just after the said relevant transfer valve has closed and so that as the said secondary valve thereof is opened, about fifty percent or more of the combustible mass has been combusted. With combustion completed and the said power piston moving towards BDC, the gas from the combustion chamber flows through the secondary port and open valve thereof to expand substantially to atmosperic before the exhaust valve of the said cylinder is opened when the piston thereof is at forty power CA degrees before BDC. This initiates the exhaust phase of the said cylinder and as the piston thereof moves towards TDC, it positively scavenges the said cylinder untill the next transfer phase thereinto begins which starts the next cycle thereof and as is described above. The operation of the other power cylinder has the same said valve and cyclic operation as that described above for the said respective power cylimder but as is obvious, it occurs 180 power CA degrees before and after it occurs in the said respective cylinder.
The alternative V configuration of FIG 7 has two units being in the said configuration with each said unit being one bank of cylinders for the said engine whilst the pumping cylinders 5 thereof are located to the inside of the said V, and of the power cylinders. A single pumping crankshaft 2 causes reciprocation of both said pumping pistons 16 whilst a single power crankshaft 1 causes reciprocation of all said power cylinders.

Claims

A two-stroke internal combustion engine comprising at least one unit having a pumping cylinder (5), a pumping piston (16) reciprocally movable in said pumping cylinder (5), two power cylinders (3, 4), a respective power piston (13, 14) reciprocally movable in each said power cylinder (3, 4), each said power cylinder (3, 4) having an associated combustion chamber (22, 25), the pumping piston (16) reciprocating at a cycle speed twice that of the power pistons (13, 14) and said power pistons (13, 14) being phased about one stroke apart, a cylinder head (19) closing the top ends of all said cylinders (3, 4, 5), said cylinder head (19) having two transfer ports (21, 24) therethrough enabling said pumping cylinder (5) to communicate with said power cylinders (3, 4), transfer valves (8, 10) controlling communication between the pumping cylinder (5) and the power cylinders (3, 4), at least two exhaust ports (23, 26) allowing exhaust gases to flow from the power cylinders (3, 4) when the associated power piston (13, 14) is at about or before its bottom dead centre position, at least one intake port (20) through the head (19) and communicating with the pumping cylinder (5), intake valve means (12) associated with the intake port (20) and allowing a major portion of an intake charge to be induced into the pumping cylinder (5) when the pumping piston (16) is moving away from its top dead centre position and said pumping piston (16) alternately transferring the intake charge into the power cylinders (3, 4) through the transfer ports (21, 24) as the pumping piston (16) moves towards its top dead centre position, said pumping piston (16) leading to the top dead centre position the power piston (13, 14) of the cylinder to which the charge is transferred, characterised in that:
the flow of the exhaust gases through the exhaust ports (23, 26) is controlled by exhaust poppet valves (9, 11);
the exhaust ports (23, 26) extend through said head;
the exhaust valves (9, 11) begin to open when the power piston (13, 14) is positioned between 90 degrees after top dead centre and 198 degrees after top dead centre and close when the power piston (13, 14) is positioned between 162 degrees before top dead centre and 10.8 degrees before top dead centre;
the transfer valves (8, 10) are poppet valves;
the transfer valves (8, 10) begin to open when the pumping piston (16) is positioned between 72 degrees after top dead centre and 288 degrees after top dead centre and close when the pumping piston (16) is positioned between 72 degrees before top dead centre and 72 degrees after top dead centre;
each exhaust valve (9, 11) closes before the associated transfer valve (8, 10) closes;
the intake valve means begin to open when the pumping piston (16) is positioned between top dead centre and 144 degrees after top dead centre and close when the pumping piston (16) is positioned between 216 degrees before top dead centre and 14.4 degrees before top dead centre; and
the pumping piston (16) leads the power piston (13, 14) to which the intake charge is to be transferred to the top dead centre position by less than 108 power piston (13, 14) degrees before top dead centre.
The engine of claim 1 wherein said power pistons (13, 14) are reciprocated by a mainshaft, said mainshaft causes reciprocation of said pumping piston (16), and the pumping cylinder (5) is spaced substantially equal distances from each said power cylinder (3, 4).
The engine of claim 2 wherein the combustion chambers (22, 25) are in constant communication with their respective said power cylinders (3, 4).
The engine of claim 2 including a respective secondary valve (27, 28) defining a constant volume combustion chamber (22, 25) between it and the associated said transfer valve (8, 10), said secondary valves (27, 28) time communication between the combustion chambers and the power cylinders (3, 4), the secondary valve (27, 28) of an associated power piston (13, 14) begins to open when said power piston (13, 14) is between 86 degrees before its top dead centre position and 94 degrees after its top dead centre position and closes when the pumping piston (16) is positioned between 288 degrees before top dead centre and its top dead centre position.
The engine of claim 4, wherein said pumping piston (16) performs substantially all of the compressive work, said transfer valve (8, 10) and the associated said secondary valve (27, 28) close about when combustion commences and the secondary valve (27, 28) closes about when the associated said transfer valve (8, 10) opens.
The engine of claim 3 wherein said transfer and exhaust valves have a valve head, said valve heads of the transfer and the exhaust valves being located at least substantially axially above the associated power cylinder (3, 4) and to one side thereof, said heads of the transfer valves (8, 10) being located higher in the cylinder head (19) than the heads of the exhaust valves, walls of the combustion chamber (22, 25) from around the transfer valves (8, 10) extending substantially towards the mainshaft so that the walls act to direct the charge from the chamber in a downward direction and said pumping piston (16) performs only part of the compressive work on the charge.
The engine of claim 2 wherein the pumping piston (16) is reciprocated by a shaft (2) driven from the mainshaft (1) and the pumping piston shaft (2) has a longitudinal axis located above a longitudinal axis of the mainshaft (1), said cylinders (3, 4, 5) have longitudinal axes parallel to one another.
The engine of claim 2 wherein the pumping piston (16) is reciprocated by a shaft (2) driven from the mainshaft (1), said pumping piston shaft (2) including drive means (46) for operating said valves or other engine auxiliary device.
The engine of claim 2 wherein that portion of the mainshaft between said power cylinders (13, 14) includes means for driving valves or other engine auxiliary device.
The engine of claim 1 wherein said pumping cylinder (5) is located within the engine at a higher location than said power cylinders (3, 4).
The engine of claim 1 including two or more said units arranged in a V configuration with all said power pistons (13, 14) being reciprocated by a common said mainshaft (1) and said pumping piston (16) being reciprocated by a separate shaft (2).
The engine of claim 1 including a crankcase (47), transfer ports (34) in a lower portion of the pumping cylinder (5) for communicating with the crankcase (47), said transfer ports (34) in said pumping cylinder (5) being uncovered when said pumping piston (16) is near its bottom dead centre position and crankcase intake valve means (32) timing the communication between a crankcase intake port (33) and said crankcase (47) so that a charge is induced while the pumping piston (16) is moving towards its top dead centre position.
The engine of claim 2 wherein the pumping cylinder (5) is positioned between the power cylinders (3, 4) and the distance between said power cylinders (3, 4) is less than the sum of the pumping cylinder (5) bore and two wall thicknesses separating the pumping cylinder (5) and one said power cylinder (3, 4), the engine further including a turbocharger (41) coupled to an exhaust manifold (23) of each said power cylinder (3, 4).
The engine of claim 13 including a pressurised intake manifold (42) leading from the turbocharger (41) and communicating with the pumping cylinder intake port (20) and wherein said crankcase intake port (33) is naturally aspirated.
The engine of claim 2 wherein said intake valve means (12) closes when said pumping piston (16) is positioned between 108 degrees before top dead centre and 72 degrees before top dead centre and said transfer valves (8, 10) open when the pumping piston (16) is positioned between 108 degrees after top dead centre and 288 degrees after top dead centre.
The engine of claim 2 including a variable valve timing mechanism.
The engine of claim 6 wherein the walls of the combustion chamber (22, 25) which direct the charge downwardly define a major portion of the volume of the combustion chamber (22, 25).