GB2144801A

GB2144801A - Reciprocating-piston fuel-injection internal-combustion engine

Info

Publication number: GB2144801A
Application number: GB08419248A
Authority: GB
Inventors: Norman David Griffiths
Original assignee: Individual
Current assignee: Individual
Priority date: 1983-07-29
Filing date: 1984-07-27
Publication date: 1985-03-13
Also published as: GB8419248D0; GB2144801B; GB8320528D0

Abstract

The engine has a reciprocating compressor (1,2) that supplies air to a power/exhaust cylinder and piston (10, 11) through e.g. valves in the cylinder-head (4). The compressor may also feed compressed-air through a passageway (9) to form a layer of clean air adjacent the piston crown, at or about BDC. As the piston (11) closes off the passageway (9) and approaches TDC a further charge of air at high pressure and temperature is admitted from the compressor through an inlet valve (12), Fig. 3, into the combustion chamber (13) and meets the layer of air. At or about TDC fuel is injected into the combustion chamber. Combustion may be initiated by a glowplug (15), Fig 1. The air from the inlet valve may be admitted to the combustion chamber with a swirling motion and "chased" by the injected fuel. The compressor may be driven by a common crankshaft. The air may pass from the compressor to the inlet valve through a transfer passageway (8), which may be heated by the exhaust manifold (18). Alternatively, a rotary compressor may be employed, Figs. 9 to 11. <IMAGE>

Description

SPECIFICATION Piston engine arrangement The object of this invention is to provide an internal combustion engine of the reciprocating piston kind and of the type employing fuel injection whereby several practical advantages are obtained as hereinafter referred to including efficient economic and reliable operation, smoother running and cleaner exhaust emission.

Thus the invention seeks to retain the advantages in these respects of a compression ignition or diesel engine but in which the combustion "knock" associated with such engines is avoided by employing a novel or improved ignition/combustion arrangement which also allows the optimum amount of throttling of the air intake at low power outputs to minimise input power, frictional losses and heat loss during combustion.

According to this invention a reciprocating piston engine arrangement is characterised by air compressing means adapted to supply compressed air at high pressure and temperature for introduction through inlet valve means into a combustion chamber or chambers of power/exhaust piston/cylinder means of the engine, said engine having fuel injection means for the injection of fuel into said air at high temperature and pressure in the combustion chamber or chambers for combustion therein and operation of the engine.

The air compressing means may consist of piston/cylinder means of the engine adapted solely to have a pumping action for compressing air at high pressure and temperature for introduction through the inlet valve means to the combustion chamber or chambers of the power/exhaust piston/cylinder means of the engine in driving relationship with said air compressing piston/cylinder means. Alternatively, the compressor may be of any suitable rotary type not necessarily driven directly by the crankshaft of the engine.

In accordance with a further feature of the invention the outlet of the air compressing means communicates with the inlet valve means of the power/ exhaust piston/cylinder means via passageway means which latter serves as a reservoir for said air at high pressure and temperature.

In practice typical constructions and arrangements may be as follows, reference being made to the accompanying diagrammatic drawings in which: Figures 1 and 2 are respectively a plan view and axial plane section (taken on the line A-A of Figure 1) of the relevant part of the engine, Figure 3 is a detail underside view taken in the direction of the arrow B of Figure 2, Figure 4 is a detail cross sectional view taken on the line G-G of Figure 3.

Figures 5, 6, 7 and 8 are detail cross sectional views taken on lines.C-C, D-D, E-E, F-F respectively of Figure 1.

Figure 9 is a plan view showing an alternative arrangement corresponding to Figure 3, and Figures 10 and ii are similar views showing further developments.

Like parts are referred to by the same or similar reference numerals throughout the drawings.

Referring firstly to Figures 1 to 3 and also Figures 4 to 8 a two cylinder in line arrangement in a common cylinder block 3 is shown in which the cylinder 1 having a piston 2 therein serves as a compressor for air compression purposes only and the cylinder 10 containing a piston 11 is for power operation by combustion of fuel-air mixture in its associated combustion chamber 13 in the cylinder head 4. The pistons 2 and 11 are connected to a common crankshaft (not shown) in the usual manner.

Air is admitted to the compressor cylinder 1 through an intake port 5 and via a non-return inlet valve 6 shown of the spring loaded poppet type (see Figure 7) which is preferably operated by valve actuating mechanism in any suitable manner such as by an overhead cam shaft 100 and in timed relationship with the engine crankshaft.

The cylinder 1 also has a non-return outlet valve 7 which is simply spring loaded (i.e. not cam operated - see Figure 8) for passage of air at high pressure and temperature into a transfer passageway 8, which passageway 8 communicates with the inlet valve 12 of the power cylinder 10. The transfer passageway 8 serves as a reservoir for air at high pressure and temperature resulting from compression in the cylinder 1.

The inlet valve 12 of the power cylinder 10 is shown of the spring loaded bucket type (Figure 6) and is mechanically opened by the cam shaft 100 at the desired time via lever and linkage mechanism 101. The power cylinder 10 also has a cam shaft operated exhaust valve 16 shown of the spring loaded poppet type (see Figure 5). As shown the valves 6, 12 and 16 are operated in timed relationship by the common cam shaft 100.

However, valves and valve operating mechanism of any suitable kind may be employed.

After having drawn in a charge of air through the intake port 5 and past the inlet valve 6 on a downward stroke, the compressor piston 2 then returns to compress the charge in its cylinder 1 to a high pressure and temperature. As the compressor piston 2 approaches top dead centre (TDC) the spring loaded non-return outlet valve 7 is lifted so that the charge at high pressure and temperature passes into the transfer passageway 8.

Part of the transfer passageway 8 is shown accommodated in the exhaust manifold 18 (Figure 3) to provide a heat exchanger arrangement whereby compressed air in the transfer passageway 8 is further heated by the exhaust gases. The crankshaft connection of the pistons 2, 11 is such that when the piston 2 in the compression cylinder 1 is at TDC, the piston 11 in the power cylinder 10 is at BDC and vice versa. After the non-return outlet valve 7 closes, approximately 1200 of crankshaft rotation elapses before the inlet valve 12 opens.

Therefore the air pressure in the transfer passage 8 increases due to the additional heat to which it is subjected, causing the air to accelerate toward the inlet valve 12.

The flow of air from the transfer passageway 8 to the power cylinder 10 is governed by the power cylinder inlet valve 12 which is timed to open as the power piston 11 approaches the top of the cylinder 10 allowing the compressed air to fill the combustion chamber 13. Combustion would normally commence between 5 and 300 before TDC so that each downward stroke of the piston 11 is an expansion or power stroke and each upward stroke is basically an exhaust stroke expelling the exhaust gases past the cam shaft operated exhaust valve 16. Early completion of the exhaust stroke is required (approximately 52" before TDC) to allow the fresh charge to fill the combustion chamber 13 before the next combustion can take place in a generally two stroke mode of operation.

At 52 before TDC the power piston has completed 81% of its total travel and in order to increase the effectiveness of the exhaust stroke a modified form of the Kadenacy system can be employed. For this purpose an auxiliary scavenge passage 9 (Figure 2) is shown provided in the cylinder block 3 which passageway 9 communicates an upper part of the compressor cylinder 1, i.e. adjacent the cylinder head 4 with a lower part of the power cylinder 10. The communication of the passageway 9 with the compressor cylinder 1 is positioned such that it is closed by the piston 2 before the air in the cylinder is too highly compressed and the communication of the passageway 9 with the power cylinder 10 is not open to the cylinder until the piston 11 is at the bottom of its stroke.

During the compression stroke of the piston a limited amount of medium pressure air will fill the passageway 9, but it will not be able to flow into the power cylinder 10 until the end of the expansion or power stroke of the piston 11.

The exhaust valve 16 is arranged to open early so that the exhaust gases are accelerated towards the valve and into the exhaust port 17 before the communication of the passageway 9 with the cylinder 10 is uncovered by the piston 11. This will result in the momentum of the exhaust gases drawing the air out of the passageway 9 and allowing most of the exhaust gases to leave the cylinder 10 while the piston 11 is still at the bottom of its stroke. Then, as the piston 11 moves up the cylinder 10, there will be a layer of clean air above it so that when the exhaust valve 16 closes all the exhaust gases will have been displaced from the cylinder 10 and the combustion chamber 13 is then ready to receive a further charge of compressed air through the inlet valve 12 to meet said layer of air.

The power cylinder inlet valve 12 may open as soon as the exhaust valve 16 closes and the new charge of air will rapidly fill the combustion chamber 13 since the avail able space is small and the air is highly compressed. The air is directed in on one side of the chamber 13 and flows or swirls around it. Thus waste heat is absorbed from the piston crown and chamber walls. Because it enters so rapidly the charge of air will continue to rotate after the inlet valve 12 has closed. As a result the trapped air now isolated from the compressor will flow past the fuel injector 14 (Figures 3 and 4) as fuel is sprayed into the combustion chamber 13 from the latter so that the fuel will "chase" the air giving a well dispersed and highly combustible mixture.At the commencement of fuel injection the mixture will pass to the ignitor 15 (Figures 1 and 5) before ignition takes place but, after it has commenced, the fuel will be sprayed into air which will then mix with the already burning mixture. As the mixture will be rotating or swirling within the combustion chamber 13 during combustion, there will be some turbulence helping to mix the air and fuel for full combustion after any flame front burning.

The main charge of air (e.g. 80%) may be introduced by the passageway 9 with the remainder admitted by the inlet valve 12.

At low power output there will be a short injection period so that not all of the air mixes with the fuel giving a heterogeneous charge while at high power output there will be a longer injection period so that more of the air is utilised. At higher engine speeds the air will be rotating or swirling faster and the fuel will be sprayed in at a faster rate. The rate of combustion pressure rise is governed by the rate at which fuel is injected and not by the shape of the combustion chamber or the flame speed as in the case of a spark ignition engine. Thus overall air/fuel ratio is not important and the combustion obtained is similar to that of a diesel engine but without ignition delay and knock.

Furthermore, and since the compression ratio is not as high, the combustion pressure rise may be greater and smoother with the pressure cycle approximating to that of the Otto cycle and with the maximum cylinder pressure limited by the action of the fuel injector 14.

During initial starting and for the first few revolutions of the crankshaft there will be little compression or resistance to rotation allowing kinetic energy to build up. Depending on the volume of the transfer passageway 8, the pressure and heat will build up after a number of crankshaft revolutions and thereafter each time the compressor 1, 2 introduces a charge of air into the transfer passageway 8 a similar mass will be allowed to flow into the combustion chamber 13. If the ignitor 15 is a glow plug sufficient electrical heating of the latter for starting can take place during the delay before the full air flow is achieved. Once the engine is firing the glow plug 6 should remain sufficiently hot from one combustion to initiate the next owing to the short time between the exhaust gases leaving the combustion chamber 13 and the next ignition and also the absence of an induction stroke of cold air. However, pre-ignition cannot occur since the time at which ignition starts is controlled by the timing of the fuel injection.

The fuel injector 14 may be of the Pintle type normally used in diesel engines. The high pressure fuel pump (not shown) may be the same as for a single cylinder two-stroke diesel engine. With such diesel fuel injection equipment the engine will operate on a variety of fuels with minimum adjustment. The twin cylinder arrangement can be dynamically balanced but there will be some im balance of loads on the crankshaft. However, this is less than in a conventional four-stroke two cylinder engine and its operation will be smoother relying less on the action of a flywheel.The combustions will be evenly spaced at 360 intervals instead of the normal alternate 180 and 540 intervals and each compression stroke will take place during an expansion or power stroke unlike a normal two cylinder four-stroke engine which has one compression stroke during an expansion stroke but the next compression stroke starts 1800 after an expansion stroke has been completed. Also the single combustion chamber 13 eliminates variations between combustions in different cylinders which conventional engines experience due to manufacturing tolerances and wear.

If desired the compressor cylinder 1 and piston 2 may be of greater diameter or swept volume than the power cylinder 10 and piston 11 but of lighter construction in order to maintain dynamic balance.

This would have the effect of supercharging the power cylinder 10. Some of the excess air may be tapped off the compressor cylinder 1 for internal cooling around the exhaust valve 16.

Owing to the smoother and even running of the engine which is obtainable the construction of the engine can be lighter since stresses due to combustion and vibration are reduced. As a result the engine is suitable for use in motor vehicles of the private car type and is capable of efficient operation on low grade lead free diesel type fuels. A high compression ratio, e.g. of the order of 14:1 to 18:1 is possible.

Increased volumetric efficiency of the compressor operation of the engine is obtainable since the incoming air charge is heated to a less extent than that which occurs in the working cylinder of a conventional engine whilst there is also room for a larger inlet valve 6. The compressor cylinder 1 and piston 2 are not contaminated by combustion products and thus can be effectively lubricated for close clearance operation.

Since every down stroke of a piston 11 is a power stroke the combustion chamber 13 is maintained at a more uniform operating temperature and is not subject to cooling by cold air admission.

Thus increased thermal efficiency of the engine is obtained and which again is in the interests of reduced toxic exhaust emission.

In general the engine is capable of providing in effect the smooth response of a spark ignition petrol engine together with the efficiency and reliability of a compression ignition or diesel engine. If desired the engine can be supercharged without adversely affecting its operating characteristics.

Particularly as regards the pistons, piston rings, cylinder block, cylinder head, crankshaft, connecting rods and much of the valve gear conventional internal combustion engine components can be largely followed whilst apart from provision for initially heating the ignitors, electrical ignition equipment (i.e. coil and distributor and drive mechanism to the latter) is not required, nor is a carburettor. A simplified arrangement of fuel pump and pipe work is also possible.

Thus in several respects low cost manufacture and maintenance of the engine is possible.

In the development shown in Figure 9 an exhaust gas driven super charger or turbo-charger 19 is used in place of the heat exchanger manifold of Figure 3 for applications where continuous high speed operation of the engine is expected. The turbo-charger 19 will assist in drawing the air from the compressor cylinder 1 and boost its pressure before forcing it through the opened inlet valve 12 into the power cylinder 10.

In the two cylinder arrangement shown in Figure 10 both cylinders act as power cylinders 10 and are supplied with hot compressed air by a turbocharger 19. During starting an electric motor 24 drives the impeller of the charger 19 via a belt or chain drive 26 and through a clutch such as a dog clutch 25 shown having ratchet form teeth for override operation. Once the engine is running the dog clutch 25 is disengaged (automatically or at will) so that thereafter the impeller of the turbocharger 19 is solely driven by the pressure of exhaust gases from the engine exhaust valves 16.

In Figure 11 a two cylinder arrangement in which both cylinders act as power cylinders 10 is again shown but in which hot compressed air is supplied by a rotary compressor 22 driven by an electric motor 24. During normal operation of the engine the electrical power to drive the motor 24 is supplied by a generator 27 driven by the engine or, as shown, by an exhaust gas driven turbine 23. The compressor 22 may be either of the dynamic or displacement types. This arrangement enables very simple controls of the turbine 23 to be provided (since no waste gate is necessary) and also of the compressor 22. Thus a simple pressure operated switch may be employed to maintain constant delivery pressure.

The notable advantage of the arrangements shown in Figures 10 and 11 where each cylinder is a power cylinder 10 firing every revolution of the crankshaft is that it is in effect the equivalent of a four cylinder four stroke engine of double the cylinder capacity. Furthermore, in both of the arrangements the electric motor 24 can also serve as a starter motor for initially turning the engine crankshaft.

These arrangements of Figures 10 and 11 are also most advantageous when applied to large multi-cylinder engines where a single turbocharger or rotary compressor takes the place of a number of cylinders. Thus a V-four engine can be provided in this way which is the equivalent of a conventional V-eight engine.

The absence of a non-return valve between the charging means such as the compressor 22 and the power cylinder inlet valves 12 is advantageous in that the resulting heat loss from the supplied air before combustion reduces the input power needed to compress the air into the combustion chamber.

Claims

1. A reciprocating piston engine arrangement characterised by air compressing means adapted to supply compressed air at high pressure and temperature for introduction through inlet valve means into the or each combustion chamber of power/exhaust piston/ cylinder means of the engine, the engine having fuel injection means for the injection of fuel into said air at high pressure and temperature in the or each combustion chamber for combustion therein and operation of the engine.

2. A reciprocating piston engine arrangement characterised by air compressing means adapted to supply air under pressure into the or each cylinder of power/exhaust piston/cylinder means of the engine wherein an inlet passageway from the air compressing means to the cylinder effects introduction of a charge of such air into the cylinder as a layer of air adjacent the crown of a piston in said cylinder when the piston is at or about the bottom dead centre (BDC) position so as to follow and scavenge early outflow of exhaust gases from the cylinder on a compression stroke of the piston, and wherein separate inlet valve means is provided to the combustion chamber of the cylinder from the air compressing means for admitting a further charge of compressed air at high pressure and temperature into the combustion chamber in order to meet said abovementioned layer of air as the piston approaches the top dead centre position TDC), the engine having fuel injection means for the injection of fuel into the highly compressed high temperature air trapped in the combustion chamber when the piston is at or about the top dead centre position for combustion of the fuel with said air for effecting power stroke operation of the piston.

3. A reciprocating piston engine arrangement according to claim 2 wherein the inlet passageway is arranged to be closed at the appropriate time by the piston on early compression stroke movement of the latter.

4. A reciprocating piston engine arrangement according to claim 1, 2 or 3 wherein exhaust valve means to the cylinder is arranged to open early during the latter part of a power stroke movement of the piston to obtain early commencement of pressurised outflow of exhaust gases from the cylinder and combustion chamber.

5. A reciprocating piston engine arrangement according to any of the preceding claims wherein the combustion chamber is provided with an ignition device such as a glow plug to at least initially effect combustion of fuel with highly compressed high temperature air contained in the combustion chamber.

6. A reciprocating piston engine arrangement according to any of the preceding claims wherein the air compressing means consists of piston/cylinder means arranged to solely have a pumping action for supplying air under pressure to the power/ exhaust piston/cylinder means of the engine.

7. A reciprocating piston engine arrangement according to claim 6 wherein the piston/cylinder air compressing means forms part of the engine and is driven by the power/exhaust piston/cylinder means of the engine e.g. by a common crank-shaft.

8. A reciprocating piston engine arrangement according to claim 6 or 7 wherein the or each cylinder of the piston/ cylinder air compressing means directly communicates by the inlet passageway with a corresponding cylinder of the power exhaust piston/cylinder means of the engine.

9. A reciprocating piston engine arrangement according to claims 6, 7 or 8 wherein the piston/ cylinder means of the air compressing means is of greater swept volume than that of the power/exhaust piston /cylinder means so as to supercharge the latter.

10. A reciprocating piston engine arrangement according to any of claims 1 to 5 wherein the air compressing means consists of a rotary type compressor driven by the engine or otherwise driven e.g. turbo driven by exhaust gases from the engine.

11. A reciprocating piston engine arrangement according to any of the preceding claims wherein the further charge of compressed air at high pressure and temperature from the compressor means is admitted to the cylinder via transfer passageway means acting as a reservoir for containing said air at high pressure and temperature.

12. A reciprocating piston engine arrangement according to claim 11 wherein the transfer passageway means is at least partly included in a heat exchanger arrangement of or to the engine exhaust system whereby compressed air in the transfer passageway is further heated by the exhaust gases.

13. A reciprocating piston engine arrangement according to any of the preceding claims wherein the inlet valve means is arranged to admit the further charge of air at high pressure and temperature into the combustion chamber with a swirling motion.

14. A reciprocating piston engine arrangement according to claim 13 wherein the fuel injection means of the engine is arranged to inject fuel into the combustion chamber so that the injected fuel follows or "chases" the swirling motion of the air admitted into the combustion chamber.

15. A reciprocating piston engine arrangement substantially as herein described with reference to any of the embodiments thereof shown in the accompanying drawings.