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JP3859595B2 - Internal combustion engine having a single crankshaft and having opposing cylinders with opposing pistons - Google Patents
Internal combustion engine having a single crankshaft and having opposing cylinders with opposing pistons Download PDF
Expired - Fee Related
JP2005500450A5 (en
JP2005500450A (en
ピーター ホフボアー
ハリミ エドワード メイヤー
2000-12-15 Application filed by ハリミ エドワード メイヤー filed Critical ハリミ エドワード メイヤー
2000-12-15 Priority to PCT/US2000/034122 priority Critical patent/WO2002048524A1/en
2005-01-06 Publication of JP2005500450A5 publication Critical patent/JP2005500450A5/ja
F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
F02B75/00—Other engines
F02B75/28—Engines with two or more pistons reciprocating within same cylinder or within essentially coaxial cylinders
F02B75/282—Engines with two or more pistons reciprocating within same cylinder or within essentially coaxial cylinders the pistons having equal strokes
F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
F01B—MACHINES OR ENGINES, IN GENERAL OR OF POSITIVE-DISPLACEMENT TYPE, e.g. STEAM ENGINES
F01B7/00—Machines or engines with two or more pistons reciprocating within same cylinder or within essentially coaxial cylinders
F01B7/02—Machines or engines with two or more pistons reciprocating within same cylinder or within essentially coaxial cylinders with oppositely reciprocating pistons
F01B7/04—Machines or engines with two or more pistons reciprocating within same cylinder or within essentially coaxial cylinders with oppositely reciprocating pistons acting on same main shaft
F02B75/16—Engines characterised by number of cylinders, e.g. single-cylinder engines
F02B75/18—Multi-cylinder engines
F02B75/24—Multi-cylinder engines with cylinders arranged oppositely relative to main shaft and of "flat" type
F02B75/246—Multi-cylinder engines with cylinders arranged oppositely relative to main shaft and of "flat" type with only one crankshaft of the "pancake" type, e.g. pairs of connecting rods attached to common crankshaft bearing
F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
F16F15/00—Suppression of vibrations in systems; Means or arrangements for avoiding or reducing out-of-balance forces, e.g. due to motion
F16F15/22—Compensation of inertia forces
F16F15/24—Compensation of inertia forces of crankshaft systems by particular disposition of cranks, pistons, or the like
F02B75/02—Engines characterised by their cycles, e.g. six-stroke
F02B2075/022—Engines characterised by their cycles, e.g. six-stroke having less than six strokes per cycle
F02B2075/025—Engines characterised by their cycles, e.g. six-stroke having less than six strokes per cycle two
F02B3/00—Engines characterised by air compression and subsequent fuel addition
F02B3/06—Engines characterised by air compression and subsequent fuel addition with compression ignition
F02B37/00—Engines characterised by provision of pumps driven at least for part of the time by exhaust
F02B37/004—Engines characterised by provision of pumps driven at least for part of the time by exhaust with exhaust drives arranged in series
F02B37/04—Engines with exhaust drive and other drive of pumps, e.g. with exhaust-driven pump and mechanically-driven second pump
F02B37/10—Engines with exhaust drive and other drive of pumps, e.g. with exhaust-driven pump and mechanically-driven second pump at least one pump being alternatively or simultaneously driven by exhaust and other drive, e.g. by pressurised fluid from a reservoir or an engine-driven pump
F02B39/00—Component parts, details, or accessories relating to, driven charging or scavenging pumps, not provided for in groups F02B33/00 - F02B37/00
F02B39/02—Drives of pumps; Varying pump drive gear ratio
F02B39/08—Non-mechanical drives, e.g. fluid drives having variable gear ratio
F02B39/10—Non-mechanical drives, e.g. fluid drives having variable gear ratio electric
1. Introduction The present invention relates generally to two-stroke internal combustion engines, and more particularly to a two-stroke internal combustion engine having two opposed cylinders each having a pair of opposed pistons.
The design and manufacture of internal combustion engines for the automotive and light aircraft industries is an advanced technical field. In order to survive commercially, the new engine will save energy and raw materials (especially improved fuel consumption), environmental protection and pollution prevention, occupant safety and comfort, lower costs, without sacrificing performance. It is necessary to have a configuration in which areas such as competitive design and production methods for realizing weight reduction are greatly improved. Of these areas, any improvement at the expense of others is not commercially acceptable.
The new engine must be mechanically simple and inherently low in mechanical loss, and well-structured to increase combustion efficiency and reduce unburned gas emissions. In particular, the new engine specifically addresses the reduction of mechanical losses by eliminating many of the main causes of friction in internal combustion engines, has a combustion chamber with a volume and design suitable for optimal combustion efficiency, The configuration should be suitable for utilizing advanced supercharging and direct fuel injection technologies.
The new engine should be lighter and preferably have a reduced height profile that is more appropriate to install and increases occupant safety. In automobiles, the engine can be installed under the seat or floor area by reducing the height profile. In light aircraft, the low profile allows the engine to be mounted directly in the wing, eliminating the need for engine cowling.
New engines should be configured in a mechanically balanced manner to minimize noise and vibration. Ideally, a practically minimal engine, such as a two-cylinder type, should be sufficiently balanced, so that a large engine can be configured by connecting small engines. In the case of low loads, all parts of the engine (and incidental mechanical losses) can be separated without breaking the balance of the engine.
2. DESCRIPTION OF THE PRIOR ART Although continuous external combustion technologies such as Stirling engines and fuel cells are virtually certain to provide low exhaust and high efficiency engines for cars and light aircraft, these technologies are weight, space Because of its inherent weaknesses in driveability, engine density and cost, it will not be a promising substitute for internal combustion engines in the near future. The internal combustion piston engine will continue to be the primary power plant for these applications for the time being.
Currently, four-stroke internal combustion engines are the mainstream in the automobile market, and in-line four cylinders are common. The need for at least four cylinders to achieve the appropriate rate of power stroke has determined the size and shape of this engine and therefore the designer's how to install the engine in the car body. The selection range is also greatly limited. The small cylinders of these engines are generally not suitable for efficient combustion or reduction of unburned gas emissions. The configuration of in-line four cylinders also has a drawback in terms of passenger comfort. This is because a very unbalanced free mass force (free-mass force) is generated, resulting in high levels of noise and vibration.
Engine designers have long compared to a 4-stroke engine with a 2-stroke engine, because each cylinder has one output stroke during one revolution of the crankshaft. It has been considered very advantageous in that the number of cylinders is half. If the number of cylinders is small, the mechanical complexity and bulk of the engine is reduced. A two-stroke engine is also inherently less mechanically complex than a four-stroke engine because it can greatly simplify the intake and exhaust port opening and closing mechanisms.
However, the two-stroke engine has some drawbacks and has limited use. A two-stroke engine has a disadvantage in terms of average effective pressure compared to a four-stroke engine (that is, volume efficiency is poor). This is because a large portion of each stroke is used and eliminated from the output stroke to eliminate (scavenge) the combustion products from the previous output stroke and to replenish the combustion air. Scavenging is also inherently problematic, especially when the engine must operate at various speeds and load conditions. It is well known that the two-stroke compression ignition (diesel) engine has other disadvantages such as poor starting characteristics and a large particle emission amount.
Modern supercharging and direct fuel injection methods overcome many of the limitations previously associated with 2-stroke engines, and 2-cylinder 2-stroke engines are an effective replacement for 4-cylinder 4-stroke engines. The 2-cylinder 2-stroke engine has the same ignition frequency as the 4-cylinder 4-stroke engine. If the average effective pressure of a two-stroke engine is two-thirds that of a four-stroke engine and the effective displacement volume of each cylinder of two strokes is increased to three-half of four strokes, the output of these two engines will be the same Should be. The small but large combustion chamber of a two-stroke engine is advantageous for improving combustion efficiency and reducing unburned gas emissions. A two-stroke engine also eliminates the need for a four-stroke engine valve, making the combustion chamber design more flexible.
It is also known that current engines have many factors that cause friction losses, and reducing these friction losses can increase engine efficiency. The largest source of friction loss, which accounts for about half of the total friction loss of current automobile engines, is due to the lateral force generated by the swinging connecting rod acting on the piston. This lateral force acts on the cylinder wall. Pressing friction is generated. The magnitude of these losses is a function of the turning radius r of the crankshaft divided by the length l of the connecting rod, and this ratio is often called Σ (lambda). By reducing the Σ by increasing the length of the effective connecting rod or reducing the travel distance of the crankshaft, the maximum reduction in friction loss can be achieved.
The loss due to contact of the piston (more precisely, the piston ring) with the cylinder wall is also a function of the average speed of the piston with respect to the cylinder wall. If the piston speed can be reduced while maintaining the same output, the friction loss will be reduced.
Another important source of friction loss in current engines is the large force acting on the crankshaft main bearing. A typical in-line four-cylinder engine has five crankshaft main bearings, which are necessary to support a crankshaft that is pushed down by literally many tons of combustion force. This force must be transmitted to the engine support structure. Both the crankshaft and the engine support structure must be designed with sufficient strength (and corresponding weight) to withstand these loads.
The object of the present invention is to achieve the same performance characteristics as the current four-cylinder four-stroke engine, but the height profile is low, the weight is light, the suitability for mounting and the applicability to advanced supercharging and direct fuel injection methods are improved. Thus, it is an object of the present invention to provide a two-cylinder two-stroke internal combustion engine in which the overall mechanical balance is substantially achieved and the manufacturing cost is reduced by mechanical simplicity.
Accordingly, an engine mechanism is disclosed that utilizes a single crankshaft and two opposing cylinders having an inner end adjacent to the crankshaft. Each cylinder has opposing inner and outer pistons that are reciprocally mounted so that a combustion chamber is formed therebetween. The push rod is drivingly connected to the inner piston to the crankshaft, and the pull rod is provided to drivingly connect the outer piston to the crankshaft.
Furthermore, according to the invention, the crankshaft preferably has at least four separate crankpins for receiving the driving force from the respective pull rod and push rod. Each cylinder includes an intake port and an exhaust port formed near each end, and fuel injection means between the intake port and the exhaust port communicating with the combustion chamber.
An important feature of the present invention is that the geometry and mass of the moving parts are selected to minimize the dynamic unbalance of the engine during operation. More particularly, the effective mass of each outer piston is preferably selected such that the product of the crank radius associated with that mass is approximately equal to the product of the crank radius associated with the effective mass of each inner piston. This configuration substantially eliminates mechanical imbalance.
According to a further preferred feature of the invention, the pull rod and push rod journal of each cylinder are arranged asymmetrically so that the exhaust port of the associated cylinder opens before the intake port opens and closes before the intake port closes. With this asymmetric timing, the turbocharger can be used to improve engine efficiency.
To achieve the asymmetrical intake and exhaust port timing of the present invention while substantially maintaining the mechanical balance, an intake port is provided on the inner end adjacent to the crankshaft for one of the cylinders, and the other cylinder Is provided with an intake port at the outer end far from the crankshaft.
Yet another preferred feature of the present invention is that each of the inner pistons has a smooth end face that is convex in a plane perpendicular to the longitudinal axis of the crankshaft at the end away from the combustion chamber. The associated push rod assembly then includes a connecting rod connected to one crank pin of the crankshaft and has an outer surface that is concave and slidably engages the end surface of the inner piston. This push rod shape effectively lengthens the push rod, resulting in reduced friction loss and improved mechanical balance.
In order to receive the driving force from the outer piston of the present invention, it is preferable to provide two pull rods for each cylinder. The two pull rod assemblies are on opposite sides of the cylinder, their inner ends surround the associated crankpin of the crankshaft , and the ends remote from the crankshaft are each pivotally coupled to the distal end of the associated outer piston. .
Maximum power efficiency from the engine according to the present invention can best be achieved by supplying compressed air to the intake port of each cylinder. Accordingly, the presently preferred form of an asymmetric timing engine according to the present invention has two superchargers, each connected to the exhaust port of the associated cylinder, receiving blowdown gas from that cylinder, and its associated Supply compressed air to the intake port of the cylinder.
1. 1. Overview of the Invention As shown in FIG. 1, the engine configuration of the present invention comprises a left cylinder 100, a right cylinder 200, and a single central crankshaft 300 installed between the cylinders (for the sake of clarity, the engine support The structure is omitted in FIG. 1).
The left cylinder 100 has an outer piston 110 and an inner piston 120 having combustion surfaces 111 and 121, respectively, and a combustion chamber 150 is formed between these pistons. Similarly, the right cylinder 200 has an outer piston 210 and an inner piston 220 having combustion surfaces 211 and 221, and a combustion chamber 250 is formed between these pistons. The four pistons 110, 120, 210, and 220 are connected to separate eccentric pins (crank pins) of the crankshaft 300.
The outer piston 110 of the left cylinder is connected to the eccentric pin 311 of the crankshaft by a pull rod 411. Similarly, the outer piston 210 of the right cylinder is connected to the eccentric pin 321 of the crankshaft by a pull rod 421. Although a single pull rod is shown in FIG. 1, in the engine of the preferred embodiment, a plurality of pairs of pull rods are used, and in the figure, one cylinder is provided on the front side of the paper and one is provided on the back side of the paper. The front and back pull rods are connected to separate crankshaft bearings having the same angular position and offset geometry. Since the pull rods 411, 421 are generally always in tension during normal engine operation and need only be supported by a small compressive force at engine start, they are relatively thin and thus lightweight, as described below. Pull rods 411, 421 are in communication with the outer piston by pin means 114, 214 that pass through slots (not shown) in the cylinder wall. The outer pistons 110, 210 are long and the pins are positioned towards the rear of the piston to prevent loss of cylinder gas from the slots. The longer pull rod length relative to the crank radius reduces the friction loss in the engine.
The left cylinder inner piston 120 is connected to the crankshaft eccentric pin 312 by a push rod 412, and the right cylinder inner piston 220 is similarly connected to the crankshaft eccentric pin 322 by a push rod 422. During normal engine operation, the push rods 412, 422 are always in compression (discussed below), rather than being connected to the inner piston by pins, the push rod has concave ends 413, 423 that are connected to the inner piston. A structure is employed which is placed on the rear convex cylindrical surfaces 125, 225. This arrangement lengthens the push rod, reduces friction losses and helps balance the engine. This will be described later.
In the figure, four pistons 110, 120, 210, 220 have a plurality of piston rings, 112, 122, 212, 222, respectively, located behind the combustion surface. In an actual embodiment of the engine, along the piston body to prevent air from leaking from the port to the crankcase or from a slot (not shown) in the cylinder wall where the outer piston communicates with the pull rod. More piston rings can be used.
Each of the cylinders 100, 200 has intake, exhaust and fuel injection ports. In the left cylinder 100, the outer piston 110 opens and closes the intake port 161, and the inner piston 120 opens and closes the exhaust port 163. The fuel injection port 162 is installed near the center of the cylinder. In the right cylinder 200, the inner piston 220 opens and closes the intake port 261, and the outer piston opens and closes the exhaust port 263. As with the left cylinder, the fuel injection port 262 is installed near the center of the cylinder. The asymmetrical arrangement of exhaust and intake ports on the two cylinders helps to balance the engine dynamics. This will be described later.
The eccentric pins 311, 312, 321, 322 of the four crankshafts are respectively installed at unique positions with respect to the rotation axis 310 of the crankshaft. The eccentric pins (312 and 322) of the inner piston are farther from the rotational axis of the crankshaft than the eccentric pins (311 and 321) of the outer piston, and as a result, the moving distance of the inner piston is longer than that of the outer piston. The eccentric pin (312) of the left inner piston, which is a piston that opens and closes the exhaust ports of the two cylinders, and the eccentric pin (321) of the right outer piston are angularly advanced, and the eccentric pin (311) of the left outer piston. And the eccentric pin (322) of the right inner biston are angularly delayed (note that the rotation of the crankshaft is counterclockwise as indicated by the arrow).
The unique position of the eccentric pin contributes to both engine balance and the operation of the engine with respect to energy recovery from supercharging and exhaust jets, as described below. By balancing the engine, much of the non-rotational force on the crankshaft is canceled and the crankshaft design can be simplified. This will be described later. By using opposed pistons, the combustion capacity per cylinder is increased, and at the same time, the radius of the crankshaft is shortened, so that the height of the engine can be reduced. The push rod shape reduces the friction loss due to lateral forces on the piston while the engine is lower and more compact.
Compared to the current state-of-the-art in-line four-cylinder engine with comparable performance, the engine according to the present invention provides a substantial improvement in ease of installation, reduced friction loss and elimination of vibration. The height of an engine with opposing cylinders with opposing pistons is largely determined by the maximum range of crankshaft sweeping motion. The design with the opposed pistons reduces the crank radius to about half of the same cylinder arrangement. Therefore, the height can be reduced by about 200 mm compared to the height of the in-line 4-cylinder engine being 450 mm. The single center crankshaft and push rod configuration provides a relatively small engine with a width of about 790 mm, which is within the range of available mounting widths for automobiles. The overall volume of the engine of the present invention is reduced by approximately 40% and by weight by 30% compared to an in-line four cylinder engine.
Friction due to lateral forces acting on the piston is greatly reduced by this design. The reciprocal (Σ) of the linkage ratio of the in-line 4-cylinder engine according to the latest technology is about one third. By lengthening the pull rod and shortening the crank radius, the Σ of the outer piston of the present invention becomes about 1/12. When the connecting rod is efficiently lengthened by sliding the push rod on the convex surface of the rear part of the piston, the Σ of the inner piston is about 1/7.
The total number of pistons of the two-cylinder engine according to the present invention is the same as that of a conventional in-line four-cylinder engine. However, when the output is the same, the distance that each piston moves is short, so that the average piston speed is greatly reduced. For the inner piston, the average piston speed is reduced by about 18% compared to a typical four-cylinder engine, and for the outer piston, the average piston speed is reduced by about 39% (the length of travel is asymmetrical). Will be described later).
The opposing piston configuration substantially eliminates non-rotating combustion forces on the main bearing. This is because the pulling of the outer piston cancels out the pushing of the inner piston, and the rotational force is mainly applied to the crankshaft. Therefore, the number of main bearings can be reduced to 2, and the crankshaft and the supporting engine structure can be reduced in weight.
The engine of the present invention is basically balanced mechanically as described below, but some remaining mechanical imbalance is acceptable at the expense of asymmetrical intake and exhaust port timing. . With such an imbalance remaining, the maximum free mass force calculated for the engine is about 700 N at 4500 rpm, compared to about 10,000 N for an in-line 4-cylinder engine, a 93% reduction. Is done.
The engine configuration of the present invention is very suitable for supercharging. As shown in FIG. 1, in a preferred embodiment, each cylinder of the engine has a separate supercharger (510, 520). If there are only two cylinders, one turbocharger can be dedicated to each cylinder, which is economical and techniques such as pulse turbocharging become more practical. The supercharger is preferably an electric motor auxiliary turbocharger, which improves scavenging, improves engine performance at low speed rotation while avoiding turbo lag, and realizes energy recovery from engine exhaust. This will be described later.
2. Engine Operation FIG. 2 shows the operation of the engine according to the invention during a complete revolution of the crankshaft. FIGS. 2 (a) to 2 (h) show the piston, intake and exhaust port positions, and piston relative speed advanced by about 45 °. Note that the crankshaft in FIG. 2 is counterclockwise. The angle φ of the crankshaft is indicated by a small triangle and a dashed arrow. Since connecting rods (push rods and pull rods) intersect at various positions on the crankshaft , the four crankpins are numbered for clarity and the crankpins 1, 2, 3, 4 are respectively left outer, left inner, Connect to the right inner and right outer pistons. As an example, the slider end of the inner push rod and the convex surface of the rear of the inner piston are shown, and the "effective" length of the inner push rod is indicated by a dashed line.
FIG. 2 (a) shows the engine at a crankshaft position of 0 ° (this position is arbitrarily defined as the "top dead center" of the left cylinder, TDC). In this respect, the left outer piston (P LO ) and the left inner piston (P LI ) are very close to the maximum approach point. In the case of a direct injection engine, at approximately this rotation angle of the crankshaft, the supplied fuel is injected into the left cylinder and combustion begins (the piston surfaces on both sides of the actual engine's combustion chamber are more complex, the flatness of FIG. The piston surface is only to show the relative piston position). At this point, the intake and exhaust ports (IN and EX) of the left cylinder are completely closed by PLO and PLI , respectively. Since the timing of the piston that operates the exhaust port is about 12.5 ° earlier and the timing of the piston that operates the intake port is delayed by approximately the same amount, both pistons P LO and P LI are as indicated by the arrows. , Move slightly to the right (the left inner piston PLI has just reversed direction). Since the crank radii of the two pistons are different, the piston speeds are also slightly different.
In the right cylinder of FIG. 2A, the right inner piston (P RI ) and the right outer piston (P RO ) are close to their maximum separation point. Both the intake and exhaust ports (IN and EX) of the right cylinder are open, scavenging the exhaust gases from the previous combustion cycle ("single flow" scavenging). Like the piston in the left cylinder, both PRI and PRO have a slight velocity in this case towards the left, and the outer piston PRO has just changed direction.
In FIG. 2 (b), the pistons PLO and PLI of the left cylinder are in the output stroke and are moving away, and the outer piston has just changed its direction of movement. The inner piston moves at a much higher speed than the outer piston, as indicated by the size of the arrow. In the right cylinder, the outer piston PRO closes the exhaust port EX, and the intake port IN is still partially open due to supercharging.
In FIG. 2 (c), the left cylinder continues its power stroke, and the two pistons PLO and PLI move in opposite directions at a more equal speed. In the right cylinder, the piston PRI closes the intake port IN and the two pistons move towards each other, compressing the air in between.
In FIG. 2 (d), the left inner piston PLI opens the exhaust port EX of the left cylinder. The intake port remains closed. In this "blowdown" state, a portion of the kinetic energy of the expanded air in the combustion chamber is recovered outside for generation of turbocharging ("pulse" turbocharging) or electrical energy. In the right cylinder, the two cylinders continue the compression stroke.
In FIG. 2 (e), the left outer piston PLO opens the intake port IN, and the cylinder is scavenged. The inner piston P LI changed its moving direction. The right cylinder has reached a position corresponding to TDC, the two pistons PRI and PRO have a slight velocity towards the right, and the outer piston has just changed its direction of movement.
In FIG. 2F , the left inner piston PLI closes the exhaust port EX. The intake port IN remains open, and the cylinder is supercharged. The outer piston PLO passed its maximum travel point and reversed its direction. The right cylinder is in the power stroke and the two pistons are moving away from each other.
In FIG. 2G, the left outer piston P LO closes the intake port IN, and the two pistons P LO and P LI move toward each other and compress the air between them. The right cylinder continues its power stroke.
In FIG. 2 (h), the left cylinder continues its compression stroke and approaches the "TDC" position in FIG. 2 (a). In the right cylinder, the outer piston PRO opened the exhaust port EX. The intake port remains closed ("blowdown").
The specific angle and timing depends on the geometry of the crankshaft and the size and position of the port. The above description is only for illustrating the concept of the present invention.
3. One important goal in the design of free mass force balance engines is to balance free mass force, eliminate vibrations, and reduce cyclically changing vibration loads on crankshafts, blocks, and other structures. It is. One piston connected to the crankpin through the connecting rod generates primary (having the same frequency as the crankshaft rotation) and higher (having a frequency several times higher than the crankshaft rotation frequency) free mass force To do. In the case of the single central crankshaft configuration of the present invention having opposing cylinders with opposing pistons, the balance of both primary and higher order free mass forces is basically taken as a whole. Theoretically, each cylinder of this engine can be balanced independently, but the present invention takes a different approach, leaving some unbalance in each cylinder, which corresponds to the opposite cylinder. Offset by unbalance. This approach avoids serious design constraints that would otherwise affect engine design.
The method of mechanical balancing in the present invention is easy to understand by first examining the inherent problems associated with balancing only one cylinder. In FIG. 3, one cylinder of the engine is shown in FIG. 3 (a), and the engine balancing method according to the present invention is illustrated in FIGS. 3 (b), 3 (c) and 3 (d).
When it is assumed that the two pistons are 180 ° out of phase (that is, I 1 and I 2 are just opposite in phase as shown in FIG. 3A), the configuration of the single cylinder shown in FIG. The free mass force is balanced for the primary and secondary forces when the following two conditions are met:
r 1 is the crank radius for the inner piston,
r 2 is the crank radius for the outer piston,
l 1 is the length of the connecting rod of the inner piston,
l 2 is the length of the connecting rod of the outer piston,
m 1 is the effective mass of the inner piston,
m 2 is the effective mass of the outer piston.
However, it is difficult to satisfy both condition (1) and condition (2). This is because in an actual design, l 2 (the length of the connecting rod of the outer piston) is much larger than l 1 (the length of the connecting rod of the inner piston). The smaller the engine, the greater this difference. The same is true for the slider push rod in the preferred embodiment of the present invention where l 1 is slightly longer.
The difference in length between the two connecting rods places design constraints on both the relative travel distance of the two pistons and the relative effective mass of the pistons (when balancing the power in the cylinder). In order to satisfy the condition (1), the crank radius r 2 related to the outer piston must be made larger than the crank radius r 1 related to the inner piston at the same ratio as the length of the connecting rod. In order to satisfy the condition (2), the effective mass m 1 of the inner piston must be larger than the effective mass m 2 of the outer piston by the same ratio. Both of these requirements unduly restrict engine design. For example, as will be described below, it is desirable to increase the length of the outer piston and thus increase its mass to accommodate the second piston set. The effective mass of the outer piston also includes the contribution of the pull rod, which in the actual design is greater than the contribution of the push rod to the effective mass of the inner piston and therefore tends to make the cylinder more unbalanced. There is.
In order to avoid the limitation due to the above conditions (1) and (2), the present invention does not pursue a perfect balance of each cylinder, and the method shown in FIGS. 3 (b), 3 (c) and 3 (d) is used. We are using.
It is well understood that the free mass force of the engine configuration (ie, "V-180 °") with the basic opposing pistons of FIG. 3 (b) is balanced except for the primary force ( The higher free mass force contributed by each of the two pistons is just canceled, leaving only the primary free mass force throughout the engine). It is also known that the primary free mass force of this engine configuration is proportional to the effective mass of the piston times the crank radius. In other words, the primary free mass force is
It can be expressed.
By analogy with the engine configuration of FIG. 3 (b), it can be shown that the engine configuration of FIG. 3 (c) also has a balance of free mass forces except for the primary force. And the primary free mass force is
In order to understand how to achieve the mechanical balance, the engine configuration of the present invention is a superposition of the engines of FIGS. 3 (b) and 3 (c) as shown in FIG. 3 (d). And the sum of free mass forces is equal to:
If I 1 and I 2 are selected such that the "engine" in FIG. 3B is 180 ° out of phase with the engine in FIG. 3C, sin (I 1 + αt) = − sin (I 2 + αt ) And the primary free mass force of the "composite" engine is m 1 · r 1 −m 2 · r 2 ,
In this case, the sum of the primary free mass forces of the combined engine is zero.
Thus, the engine configuration of FIG. 3 (d) is balanced because the "engine" shown in FIGS. 3 (b) and 3 (c) is balanced except for the primary free mass force. Are balanced, and the primary free mass force of the two components "engine" is countered by setting the following:
In each engine, which is a component, one piston opens and closes the exhaust port and the other piston opens and closes the intake port. Note that the masses of the two pistons are compatible.
This balancing of the engine is particularly advantageous when the length of the connecting rod is not a decisive factor in achieving a mechanical balance. In reality, the analysis determines the effective mass of the inner and outer pistons (including the contributions of the pull and push rods) and calculates the crankshaft radii r 1 and r 2 required to achieve balance. Is relatively easy. In a preferred embodiment, increasing the effective mass of the outer piston requires that the stroke of the outer piston be much smaller than that of the inner piston, which is necessary to balance each cylinder independently. The opposite is true.
The above description assumes an engine with symmetrically timing intake and exhaust ports, two cylinders and a crankshaft. The basic configuration of the present invention having opposed cylinders of opposed pistons is basically balanced overall as described above, but in the preferred embodiment, it accepts some remaining unbalance, As will be described later, asynchronous timing of the intake and exhaust ports is enabled. Even with this remaining imbalance, the computer analysis shows that the free mass force of the preferred embodiment is on the order of less than the free mass of a standard in-line four-cylinder four-stroke engine with comparable performance. Is done.
4). Asymmetric timing of intake and exhaust ports The asymmetric timing of the intake and exhaust ports of a two-cycle engine has many important advantages. When the exhaust port opens before the intake port, the exhaust gas energy is more effectively recycled by the turbocharger, and when the exhaust port closes before the intake port, the cylinder is more effectively supercharged.
In the engine configuration of the present invention, as described above, the intake port is controlled by one piston of each cylinder, and the exhaust port is controlled by the other piston. In this configuration, not only effective scavenging ("single flow" scavenging) is performed, but also the timing of the exhaust port with respect to the intake port can be independently asymmetric.
The asymmetric timing of the two pistons of each cylinder is realized by changing the relative angular position of the corresponding crankpin (see FIG. 1). When the crankpins corresponding to the two pistons are installed 180 ° apart, the two pistons both simultaneously reach the lowest and highest positions in their respective trajectories (symmetric timing), and in a preferred embodiment of the invention, The exhaust port pin journal is angularly advanced by approximately 12.5 °, and the intake port pin journal is angularly delayed by approximately 12.5 ° (thus "top dead center" is a symmetrical timing engine. Occurs at the same crankshaft angle as in, but the two pistons have a slight common movement with respect to the cylinder). As a result, the exhaust port opens before the intake port for "blowdown" and closes before the intake port for supercharging.
As described above, in the engine configuration according to the present invention, instead of making the intake and exhaust port timings asymmetric (as described above), a certain amount of imbalance of free mass force occurs (the slight longitudinal direction of the two cylinders). Is also a cause of this imbalance, which will be discussed later). In the preferred embodiment, this imbalance is kept to a minimum by reversing the relative position of the intake and exhaust ports of a single cylinder, as shown in FIG.
FIG. 4 (a) shows a configuration of opposing cylinders having opposing pistons with symmetrical piston timing. The exhaust ports of both cylinders are in the inner configuration (ie, closest to the crankshaft) and the intake ports are in the outer configuration. The free mass force of this engine is basically balanced as described above.
FIG. 4B shows the same engine configuration with asymmetric exhaust and intake port timing. The two "engines" described with reference to FIGS. 3 (b) and 3 (c) are not out of phase in FIG. 4 (b), so this engine also has some primary free mass force that remains unerased. However, this is an engine configuration that can survive in the future because the free mass force that is not canceled is much smaller than in the case of a conventional in-line four-cylinder engine.
In a preferred embodiment, as shown in FIGS. 4 (c) and 4 (d), an optimal balance over that shown in FIG. 4 (b) is achieved by reversing the intake and exhaust ports for one of the two cylinders. Is realized. FIG. 4 (c) shows a symmetrical timing engine with the exhaust and intake ports reversed for one cylinder. If the mass of the piston is the same, this engine has the same free mass balance as the engine of FIG. FIG. 4 (d) shows the engine of the preferred embodiment. In order to reverse the position of the exhaust and intake ports of one cylinder, it is necessary to "split" the crankshaft throw to maintain the correct port timing. The free mass force of this engine is unbalanced, but these are negligible because they are less than one tenth of the secondary free mass force found in inline 4-cylinder engines. The improvement in balance is due to each inner piston being substantially 180 degrees out of phase with the outer piston of the opposite cylinder. Again, this asymmetric configuration is perfectly balanced if the inner piston lambda (crank radius divided by the connecting rod length) is equal to the outer piston lambda (to reduce friction loss further, as described below). Ignore the slight imbalance that is introduced). Thus, in a preferred embodiment, increasing the effective length of the inner piston push rod contributes to mechanical balance.
In order to achieve a mechanical balance, it is preferable to increase the effective length of the inner push rod (by increasing the radius of curvature of the cylindrical convex surface of the rear part of the inner piston). Limited by element. First, if the radius is too large, the lateral force applied to the slider will be insufficient to move the slider correctly along the convex surface. Second, if the push rod is too long, mechanical interference occurs between the push rod and the cylinder wall. Since it is also desirable to make the engine as small as possible, this second element is the limiting element in the preferred embodiment.
5). More Detailed Description of Asymmetric Timing in the Preferred Embodiment The operation of the preferred embodiment is further illustrated in FIG. 5, which shows the piston surface in the cylinder with respect to the angle of the crankshaft during a complete revolution of the crankshaft. Indicates the position. The location of the cylinder wall intake and exhaust ports is also shown. In FIG. 5, the asymmetric timing of the two pistons in each cylinder is clearly observed, the two pistons reaching the highest position during the stroke at different crankshaft angles and moving together with respect to the cylinder at "TDC". It can also be seen from the difference in crank radius that the inner piston moves longer than the outer piston. Since the intake port is actuated by the left outer and right inner pistons and the exhaust port is actuated by the left inner and right outer pistons, the intake and exhaust port dimensions for the two cylinders are slightly different.
6). Adaptability of configuration with opposed cylinders with opposed pistons when compared to larger engines In many engine configurations, the balance contributes 4, 6, 8 or more cylinders with individual pistons contributing It is taken by arranging so that the free mass force to be canceled. Often reversing weights are used, which further complicates the engine. An advantage of the present invention is that a substantially overall balance is achieved in a compact engine using only two cylinders. A larger engine can be constructed by arranging a plurality of small engines side by side and connecting their crankshafts. For the connection, means such as an electric clutch can be used, and the connection of the pair of cylinders can be disconnected when unnecessary at low load. Some engines currently use only a portion of the cylinders while running at a partial load, but the cylinder remains connected to the crankshaft and the piston continues to move within the cylinder, so the friction load on the engine continues.
1. Physical Description The presently preferred embodiment of the present invention is further illustrated in the front, top, and front cross-sectional views of FIGS. 6, 7, and 8, respectively. These drawings show an engine with a crankshaft angle of 270 ° from the TDC of the left cylinder. The engine includes a left cylinder 1100, a right cylinder 1200, and a single central crankshaft 1300 between the cylinders (engine support structure is not shown).
As shown in FIG. 8, the left cylinder 1100 has an outer piston 1110 and an inner piston 1120 that have combustion surfaces 1111 and 1121, respectively, with the two pistons forming a combustion chamber 1150 therebetween. The right cylinder 1200 similarly has an outer piston 1210, an inner piston 1220 and a combustion chamber 1250 having combustion surfaces 1211 and 1221, respectively. Each of the four pistons 1110, 1120, 1210, 1220 is connected to another eccentric pin on the crankshaft 1300.
As best seen in FIG. 7, the outer piston 1110 of the left cylinder is connected to the crankshaft by two pull rods 1411 located on each side of the cylinder, and the outer piston 1210 of the right cylinder is likewise two It is connected to the crankshaft by a pull rod 1421. Pull rods 1411 and 1421 communicate with the outer piston by pins 1114 and 1214 passing through cylinder wall slots 1115 and 1215 (see FIG. 6).
The left cylinder inner piston 1120 is connected to the crankshaft by a push rod 1412, and the right cylinder inner piston 1220 is similarly connected to the crankshaft by a push rod 1422. The push rod has concave ends 1413, 1423 that ride on the convex cylindrical surfaces 1125, 1225 at the rear of the inner piston.
Each of the four pistons, 1110, 1120, 1210, 1220 has a plurality of piston rings 1112, 1122, 1212, 1222, which are installed behind the combustion surface and along the piston body, with gas from the port to the crankcase. Or prevent the outer piston from escaping from a slot in the cylinder wall that communicates with the pull rod.
The cylinders 1100 and 1200 have intake, exhaust, and fuel injection ports, respectively. Each intake and exhaust port has a row of ports surrounding the cylinder. In the preferred embodiment, the intake port is comprised of two rows of ports (left cylinder 1161a, 1161b and right cylinder 1261a, 1261b) for improved scavenging. This will be described later. In the left cylinder 1100, the outer piston 1110 opens and closes the intake port, and the inner piston 1120 opens and closes the exhaust port 1163. The fuel injection port 1162 is near the center of the cylinder. In the right cylinder 1200, the inner piston 1220 opens and closes the intake ports 1261a and 1261b, and the outer piston opens and closes the exhaust port. Again, the fuel injection port 1262 is near the center of the cylinder.
The preferred embodiment utilizes two superchargers (1510, 1520), one for each cylinder. The supercharger is an electric motor or a generator auxiliary turbocharger. By using different turbochargers for the two cylinders, pulse turbocharging becomes practical, as will be described later.
As seen in FIGS. 6 and 8, the left and right cylinders (1100 and 1200, respectively) of the preferred embodiment have a slight vertical offset or misalignment with each other, with the left cylinder somewhat higher than the right cylinder. . According to computer analysis, this slight misalignment (on the order of 10 mm in the preferred embodiment) reduces the overall friction loss of the engine to some extent. Computer analysis further, if the offset is properly selected, will produce a small mechanical imbalance in the engine's residual imbalance, generally of opposite polarity, which will substantially reduce the engine's residual imbalance. Shows that it helps to counteract
2. Intake and Exhaust Port Timing and Crankshaft Parameters FIG. 9 in conjunction with FIG. 8 illustrates the intake and exhaust port timing of the preferred embodiment of the present invention. For illustration purposes, a crankshaft angle of 0 ° is arbitrarily defined as the top dead center (TDC) of the left cylinder. Note that TDC is here defined as the point where the two pistons in the cylinder are closest to each other. Since the timing of one piston is ahead and the other is delayed, the two pistons have a slight common speed at this point with respect to the cylinder (to the right in the figure for both cylinders).
As mentioned above, the piston located inside each cylinder is not fixed to the corresponding connecting rod by a pin, but rather, it strikes the concave cylindrical surface at the end of the rod through the crosshead slipper. (For example, the effect of reducing the lateral force on the piston and reducing friction).
For the sake of clarity, the engine is shown in FIG. 8 with a crankshaft rotation angle of 270 ° and the same crankshaft angle as in FIG. At this angle, the left cylinder pistons are approaching each other, closing all intake and exhaust ports and compressing the air in between. The right cylinder is in the power stroke and the exhaust port is not yet open.
The timing for the left cylinder is shown in FIG. 9 (a), and the right cylinder is shown in FIG. 9 (b). For the left cylinder, starting from the position shown in FIG. 8 and progressing to the end of one cycle, the timing is as follows.
As the crankshaft approaches 0 °, the gap between the inner and outer pistons narrows and the air between the pistons is compressed and heated. Near the TDC (crankshaft angle of 0 °), the outer periphery of the piston contacts, creating a "squish" area that generates a powerful flow in the combustion chamber as described below. At some point before TDC, fuel is injected from port 1162 into the combustion chamber and combustion begins.
The power stroke exceeds the crankshaft angle of 90 ° and the gap between the inner and outer pistons widens. At the time of EX OPEN, the piston 1120 arranged inside starts to open the exhaust port 1163. During the period indicated by [B] (in "Blowdown"), the kinetic energy of the expanding gas is used for pulse turbocharging. This will be described later.
When IN A OPEN, the outer piston 1110 begins to open the first row of intake or scavenging ports, 1161a. The first row of ports is arranged so that the air enters tangentially to the cylinder, generates a swirl in the cylinder, and the exhaust gas in the cylinder is scavenged through the exhaust port. These ports and the port of 1161b are inclined toward the outer end of the cylinder (about 23 ° in the preferred embodiment) so that the intake flow is tangential to the toroidal squish band of the outer piston. It has become. In FIG. 9A, scavenging is indicated by [S].
When IN B OPEN, the second column 1161b of intake or scavenging ports opens. This row of ports directs air toward the center of the cylinder, not tangential to the periphery of the cylinder edge. Air entering from the port 1161b passes through the surface of the outer piston 1110 and is directed by the central apex of the piston through the center of the combustion chamber. This scavenges the central vortex of the exhaust gas created by the swirl in the first row of scavenging ports.
Due to the asymmetric timing of the two pistons, no point in the cycle corresponds exactly to what is commonly referred to as bottom dead center (BDC). At the point B1, the inner piston reaches its maximum travel position and reverses direction. At B2, both pistons move in the same direction at the same speed (as opposed to the previously defined "TDC"). At B3, the piston of the outer arrangement reaches its maximum movement position and reverses direction.
In EX CLOSE, the piston 1120 disposed inside covers the exhaust port 1163. The cylinder is filled with compressed air using a turbocharger or supercharger until the piston located outward from EX CLOSE covers the first row of intake ports with IN A CLOSE. This will be described later. In FIG. 9A, the filling period is indicated by [C]. By closing the exhaust port before the intake port, not only can the engine be supercharged, but the amount of air entering the combustion chamber can be limited depending on the situation. For example, when the engine load is low, the amount of air entering the combustion chamber is reduced, and the amount of fuel injected correspondingly is reduced, thereby improving the travel distance and reducing exhaust. As described below, a turbocharger integrated with a motor / generator is suitable for this purpose.
The timing of the right cylinder shown in FIG. 9B is basically the same as that of the left cylinder, but is 180 ° out of phase with the left cylinder, and the functions of the pistons arranged inside and outside are opposite.
3. Crankshaft Design FIG. 10 further illustrates a crankshaft that is currently advantageous. The four crankshaft eccentric pins (crankpins) 1311, 1312, 1321, 1322 are in a unique position with respect to the rotation shaft 1310 of the crankshaft. The eccentric pins (1312, 1322) of the inner piston are further away from the rotation axis of the crankshaft than the eccentric pins (1311, 1321) of the outer piston, and the inner piston has a longer moving distance than the outer piston. As shown in the sectional views BB, CC, DD, and EE, the eccentric pin (1312) of the left inner piston, which is a piston that opens and closes the exhaust ports of the two cylinders, and the right outer piston The eccentric pin (1321) of the left outer piston is angularly advanced, and the eccentric pin (1311) of the left outer piston and the eccentric pin (1322) of the right inner piston are angularly delayed.
FIG. 11 shows the actual geometry of the preferred embodiment crankpin . The crank radius of the inner piston is 36.25 mm, and the crank radius of the outer piston is 27.25 mm. The pins associated with the pistons controlling the exhaust ports of the left and right cylinders are respectively preceded by 7.5 ° and 13.7 ° (again, the rotation of the crankshaft is counterclockwise). The pins associated with the pistons that control the intake ports of the left and right cylinders are delayed by 17.5 ° and 11.3 °, respectively. The difference between the left and right cylinder angles results from engine asymmetry, including the 10 mm vertical offset of the two pistons, as described above.
The main role of the crankshaft is to convert the reciprocating motion of the piston transmitted by the pull rod and push rod into rotational motion. When an unbalanced force is applied to the crankshaft, friction between the crankshaft and a bearing supporting the crankshaft increases. The presence of an unbalanced force complicates the engine design because the force must be mechanically transferred to an engine support structure that is strong enough to withstand it. For example, in a standard in-line four-cylinder engine, the force from all four pistons works in the same direction relative to the crankshaft and literally many tons of pressure must be transferred from the crankshaft main bearing to the engine structure. . A typical in-line four-cylinder engine has five main bearings that support the crankshaft.
With the engine configuration of the present invention, the reaction force of the inner and outer pistons of each cylinder can be substantially canceled, so that a simpler crankshaft design can be achieved. For the left cylinder shown in FIG. 4 (d), the compression and combustion forces on the two pistons are substantially equal and in opposite directions, so the pull force of the outer piston is approximately the same as the push rod of the inner piston pushes. It can be seen that it pulls against the crankshaft. As a result, there is very little torque on the crankshaft that is generated by an uncancelled lateral or vertical force (due to the slight angle difference between the pull rod and push rod and the asymmetric timing of the two pistons). appear. Therefore, the load on the main bearing of the crankshaft is very small and the central main bearing is not required, so the friction loss is much smaller than in-line four-cylinder engines of comparable performance.
4). Supercharging the Preferred Embodiment A method for supercharging the preferred embodiment is shown in FIG. FIG. 12 (a) shows conventional turbocharging and FIG. 12 (b) shows a preferred embodiment electric motor / generator auxiliary turbocharging. In the engine configuration of the present invention, two cylinders are installed at a wide interval, and the timings of the intake and exhaust ports are independent, thereby controlling the air during scavenging and intake, and energy from the exhaust gas. An important opportunity to recover. In particular, since there are only two cylinders, it is economically feasible to provide a separate turbocharger for each cylinder to enable pulsed turbocharging. In addition, if the turbocharger is equipped with an electric motor / generator, significant performance advantages are also obtained.
As is often seen in the past, the success or failure of a two-stroke design is largely determined by its scavenging ability. In order to achieve complete combustion especially for controlling the EGR rate necessary for reducing NOx, it is necessary to perform scavenging that is optimal for the entire engine.
Aggressive boost systems such as turbocharging may be necessary to ensure efficient operation at all load conditions. Instead of or in addition to a turbocharger, a belt-type supercharger can be used to meet combustion requirements. However, using both has disadvantages in terms of weight and cost. Furthermore, the supercharger may not provide sufficient combustion pressure at start-up. If only the engine efficiency is to be improved, a turbocharger according to the prior art will be sufficient. However, at start-up and under low load conditions, the exhaust energy cannot provide the positive boost necessary to support compression.
Thus, as will be described later, an electric auxiliary turbocharger would be advantageous. An electric auxiliary turbocharger such as a conventional turbocharger has an intake impeller that compresses intake air such as ambient air. This compressed air is supplied to the intake port of the cylinder. Further, similarly to the conventional turbocharger, the exhaust impeller is connected to the intake impeller. This connection causes the exhaust gas to move the intake impeller. However, since the electric auxiliary turbocharger has an electric motor, the intake impeller can be operated regardless of the exhaust impeller and regardless of the engine operating speed. Other advantages will be described below.
4 (a). In order to give a 2-stroke engine with good pressure control to a horsepower equal to or higher than that of a 4-stroke engine, it is necessary to use supercharging scavenging. Scavenging relies on an optimal pressure ratio between charge pressure and exhaust gas back pressure. The pressure ratio is mainly adapted to the engine speed and must be increased with increasing engine speed. The pressure ratio must also be adapted to the load and transient operating conditions.
This can be achieved with an electrically assisted turbocharger having a permanent magnet brushless DC motor, and electrical control of the turbo speed and hence the supply pressure can be used. Alternatively, the motor may be of an induction motor, switching magnetoresistor, or other type that generates sufficient force to perform the required function. This electrical unit is used to assist the compressor when the exhaust energy is not sufficient to compress the intake air as desired. Furthermore, the boost pressure of the engine can be dynamically controlled by the electric auxiliary turbo. Also, the size and weight of the electric auxiliary turbocharger can be much smaller and lighter than the conventional mechanically driven supercharger.
One particular advantage of an electrically assisted turbocharger connected to a two-stroke internal combustion engine with opposing pistons and opposing cylinders is an improved start, in particular a cold start. In a conventional two-stroke engine, a glow plug and a high compression ratio are required for a reliable cold start. High compression ratios generate significant friction when the engine operates at high load conditions. The electrically assisted turbocharger according to the present invention solves this problem by using an electric motor that drives the compressor before the engine is started. For example, cool air is compressed to about 3.5 psi. For example, ambient air at about −20 ° C. is supplied to the combustion chamber at about 80 ° C. Such air heating is the result of the compression process. The electric motor turns the turbocharger's pressure impeller, feeding compressed air into the cylinder and naturally heating the air.
Thus, in order to improve the cold start of the engine, the electric auxiliary turbo can heat the intake air before the ignition sequence. In a compressed engine, the ignition sequence can be delayed by controlling the fuel release timing.
4 (b). Pulsed turbocharging internal combustion engines are inherently unstable pulsating devices. The turbine can be designed to accept such an unstable flow, but if the flow is stable, it will operate more efficiently. In practice, two methods are generally used to recycle some of the available exhaust energy. Constant pressure turbocharging and pulse turbocharging. For constant pressure turbocharging, an exhaust manifold with sufficient capacity to attenuate mass flow and pressure pulses is used to essentially stabilize the flow to the turbine. The disadvantage of this method is that it does not fully utilize the high kinetic energy of the gas exiting the exhaust port, and the loss due to mixing such a high speed gas with a large amount of low speed gas cannot be repaired. In pulse turbocharging, short and thin pipes are connected from each exhaust port to the turbine to take advantage of much of the kinetic energy associated with exhaust blowdown. By appropriately grouping different cylinder exhaust ports, exhaust pulses are continuous, and overlaps are reduced as much as possible, flow instability can be kept at an acceptable level. The turbine must be specifically designed so that such pulsating flow can achieve adequate efficiency. The combination of increased available energy in the turbine and reasonable turbine efficiency, pulse systems are more commonly used in large diesel. In the case of an automobile engine, a constant pressure turbocharger is used.
Many turbocharged heavy duty engines use a split exhaust manifold system connected to a split turbine swirl casing. For example, a six cylinder engine typically uses an exhaust manifold consisting of two branches. One branch covers the exhaust ports of the cylinders 1, 2, 3 and the other covers the cylinders 4, 5, 6. In the case of the standard ignition sequence of 1-5-3-6-2-4, the exhaust pulses from the cylinder are emitted alternately from one branch and the other branch, with a 120 ° crank between each exhaust pulse. A corner is formed. The flow path of the exhaust gas remains divided from the manifold branch to the peripheral outlet of the turbine wheel through the spiral by the turbine of the divided casing. In this way, the split manifold system prevents the blowdown pulse of each cylinder from interfering with the discharge of gas from the previously ignited cylinder.
Unfortunately, the high gas velocity generated when the exhaust valve is opened is essentially lost when the pulse exits the exhaust port, enters the manifold and hits a large area of the exhaust port on the way to the turbine casing inlet . As a result, the turbocharger turbine casing is designed with a converging nozzle section to recreate the high speed necessary to drive the turbine wheel. The exhaust gas must flow from the relatively small flow area at the throat portion of the nozzle section, creating high back pressure in the manifold branch, which increases engine pumping loss.
In the case of the engine according to the invention, the speed generated by the cylinder blowdown process can be used to drive the turbine directly. Since the exhaust gas enters the turbine casing immediately after leaving the collection chamber of the cylinder, there is no need to use a nozzle section in the turbine casing. In addition, because one turbocharger is used per cylinder, the turbine casing does not need to be divided internally, and the exhaust gas can enter completely without being divided up to the periphery of the turbine wheel, thereby improving turbine efficiency. Become the maximum.
The speed of the blowdown exhaust gas from the cylinder to the turbine wheel can be maintained by the unique design of the engine according to the invention and the use of one turbocharger per cylinder. Since there is no nozzle section in the turbine casing, the back pressure in the exhaust system when the piston exhausts cylinder air is very low. Unlike standard split manifold systems, the engine of the present invention has a much larger pressure differential within the cylinder. As a result, the fuel consumption rate is greatly improved compared to a standard turbocharged two-cycle or four-cycle engine.
4 (c). A high-efficiency cylinder scavenging suitable for single-flow scavenging requires a clear frontline between the intake and exhaust gases.
Widely used loop scavenging or backflow scavenging cannot meet current and future needs for light aircraft and automotive engines. This is because exhaust gas and intake air are mixed. Among possible single-flow scavenging schemes, poppet exhaust valves, opposing pistons, or split single design, that of opposing pistons can achieve the highest level of scavenging efficiency and the least exhaust gas and fresh intake air mixing The port configuration is the most promising.
5). Push rod and pull rod design Approximately 50% of the total friction loss of the engine is due to the lateral force generated by the rotating connecting rod acting on the piston, ie pushing the piston against the cylinder wall. The lateral force generated by the short connecting rod is large and the lateral force generated by the long connecting rod is small (an infinitely long connecting rod does not exert any lateral force on the piston, but this also , Infinitely big and infinitely heavy). It is desirable to reduce these lateral forces, i.e. friction losses, without increasing the size or weight of the connecting rod.
Only the compression load is applied to the connecting rod of the inner piston of the engine according to the present invention, and no wrist pin is required. This can be replaced by a concave radius having a large diameter, and a crosshead slipper that slides on it strikes and the connecting rod slides thereon (see FIG. 13). For this design to be effective, the force at the end of the crosshead slipper must be greater than zero. This is true when the coefficient of friction between the crosshead slipper and the sliding of the connecting rod is less than 0.45. With this configuration, the theoretical rod length is increased by more than 100 mm, and the lateral direction applied to the piston and the friction loss in the engine are reduced. Furthermore, since the Σ of the piston disposed inside becomes small, the above-mentioned free mass force is also reduced.
The outer piston transmits the reciprocating motion to the crankshaft through the two connecting rods outside the cylinder (see FIG. 14). These connecting rods are only tensioned and are therefore called pull rods. Again, if the length of the pull rod is long, the friction is greatly reduced. The pull rod can be reduced in weight by utilizing a constant tension of a buckling load and designing it to be long and thin.
6). The purpose of the combustion chamber design combustion system is as follows.
1. Reduce fuel consumption by optimal thermodynamic process.
2. Reduce pollutants in exhaust gas by optimizing reduction kinetics.
3. Increase output.
4). Reduce powertrain noise and stress.
In terms of fuel consumption, the working gas temperature can be much higher than the wall temperature, so in an internal combustion engine, the periodic combustion stroke is superior to the continuous combustion stroke (gas turbine, Stirling engine, etc.). This greatly increases the thermodynamic efficiency. Among periodic internal combustion engines, DI diesel is the most promising in that optimum heat generation is possible by controlling the injection amount with respect to the crank angle. In order to achieve a desirable combustion stroke (optimal heat generation), a combination of the correct injection amount and swirl characteristics is required.
For the reduction of pollutants, the engine according to the invention has great potential. Since this engine has no valves, the shape of the combustion chamber can be designed completely freely. An example is shown in FIG. In this, the states immediately before the top dead center of the combustion chamber (FIG. 15A), top dead center (FIG. 15B), and immediately after top dead center (FIG. 15C) are shown.
The combustion chamber is formed by an exhaust piston having a toroidal shape that fits an intake side piston having a reverse profile. The piston forms a large area squish band, which creates a high strength swirl near top dead center. This conventional combustion system, provided by the opposed piston design, has the potential to improve exhaust, fuel consumption, power and comfort.
In addition to the features found in conventional combustion systems, the engine according to the present invention can provide a novel combustion system as shown in FIGS. 16 (a) and 16 (b). By dividing the capacity of the cylinder into the combustion chamber and the cylinder, it is possible to provide a NOx reduction heat sink or a catalytic converter between the combustion chamber and the cylinder (see FIG. 16A). For reaction kinetic reasons, and in order to maintain an optimal configuration for scavenging, the converter is attached to the exhaust piston. Fuel is injected directly into the combustion chamber by injection. Such a combustion system can be a breakthrough to enable combustion of very low exhaust pollutants without sacrificing fuel consumption, power or comfort.
FIG. 16B shows a combustion chamber design that has a spherical shape at a position very close to the fuel injection means, maintains a high pressure during fuel injection, does not require a narrow channel, and avoids problems associated with the narrow channel. Indicates.
In the foregoing , specific embodiments of the present invention have been described in detail. It will be appreciated that deviations from the disclosed embodiments are also within the scope of the invention and can be clearly modified by those skilled in the art. Applicants intend that the present invention also include other embodiments known in the art that perform the same functions as those disclosed. This specification should not be construed to unduly narrow the entire protection scope of the present invention.
All means or steps, functional element structures, materials, acts, or equivalents in the following claims are also intended to function in combination with other specifically claimed elements. Shall be included.
1 is a schematic diagram of an engine configuration of the present invention. It is a figure which shows the basic operation cycle of the engine of this invention, and shows the piston position in the cylinder in each rotation angle of a crankshaft. It is a figure which shows the basic operation cycle of the engine of this invention, and shows the piston position in the cylinder in each rotation angle of a crankshaft. It is a figure which shows the basic operation cycle of the engine of this invention, and shows the piston position in the cylinder in each rotation angle of a crankshaft. It is a figure which shows the basic operation cycle of the engine of this invention, and shows the piston position in the cylinder in each rotation angle of a crankshaft. It is a figure which shows the basic operation cycle of the engine of this invention, and shows the piston position in the cylinder in each rotation angle of a crankshaft. It is a figure which shows the basic operation cycle of the engine of this invention, and shows the piston position in the cylinder in each rotation angle of a crankshaft. It is a figure which shows the basic operation cycle of the engine of this invention, and shows the piston position in the cylinder in each rotation angle of a crankshaft. It is a figure which shows the basic operation cycle of the engine of this invention, and shows the piston position in the cylinder in each rotation angle of a crankshaft. It is a figure which shows the method for implement | achieving the balance of the free mass force in an engine. It is a figure which shows the method for implement | achieving the balance of the free mass force in an engine. It is a figure which shows the method for implement | achieving the balance of the free mass force in an engine. It is a figure which shows the method for implement | achieving the balance of the free mass force in an engine. It is a figure which shows that it is necessary to install the intake and exhaust ports in two cylinders in different positions in order to make the timing of the intake and exhaust ports of the engine asymmetric. It is a figure which shows that it is necessary to install the intake and exhaust ports in two cylinders in different positions in order to make the timing of the intake and exhaust ports of the engine asymmetric. It is a figure which shows that it is necessary to install the intake and exhaust ports in two cylinders in different positions in order to make the timing of the intake and exhaust ports of the engine asymmetric. It is a figure which shows that it is necessary to install the intake and exhaust ports in two cylinders in different positions in order to make the timing of the intake and exhaust ports of the engine asymmetric. FIG. 6 is another diagram illustrating the asymmetric timing of the preferred embodiment, for the left cylinder [5 (a)] and the right cylinder [5 (b)], a linear plot of piston position during one revolution of the crankshaft. Is. 1 is a front view of a preferred embodiment of the present invention. 1 is a top view of a preferred embodiment of the present invention. It is front sectional drawing by AA of preferable embodiment of this invention shown in FIG. FIG. 4 shows the detailed timing of a preferred embodiment of the present invention, showing the opening and closing of the intake and exhaust ports of two cylinders according to the angle of the crankshaft. FIG. 4 shows the detailed timing of a preferred embodiment of the present invention, showing the opening and closing of the intake and exhaust ports of two cylinders according to the angle of the crankshaft. They are the side view of the crankshaft of preferable embodiment, and sectional drawing 10 (a) -10 (d) of a crankpin part. FIG. 6 is a schematic diagram of the crankpin geometry, showing the relationship between engine balance, asymmetric timing and crankshaft design. It is the schematic which shows the conventional supercharging. 1 is a schematic diagram illustrating supercharging in a preferred embodiment. FIG. 3 is a detailed view of a preferred embodiment push rod. It is a detail drawing of the pull rod of a preferred embodiment. FIG. 2 is a detailed view of a combustion chamber of a preferred embodiment. FIG. 2 is a detailed view of a combustion chamber of a preferred embodiment. FIG. 2 is a detailed view of a combustion chamber of a preferred embodiment. It is a figure which shows the example of a design of another combustion chamber. It is a figure which shows the example of a design of another combustion chamber.
An internal combustion engine,
Two opposing cylinders, each cylinder having two opposing pistons, an intake port, and an exhaust port; and
A crankshaft having a crankpin connected by a piston and a push rod or a pull rod;
A turbocharger that is provided for each cylinder, receives blowdown gas from the exhaust port, and supplies compressed gas to the intake port;
Due to the arrangement of the crankpin, in two cylinders, the piston is operated so that the exhaust port is opened before the intake port is opened and the exhaust port is closed before the intake port is closed.
Internal combustion engine.
The internal combustion engine according to claim 1,
Each turbocharger further includes an electric auxiliary motor for allowing the compressed gas supplied to the intake port to operate regardless of blowdown gas from the exhaust port and regardless of engine operating speed. An internal combustion engine comprising:
Each turbocharger further includes an electric auxiliary motor for allowing the compressed gas supplied to the intake port to operate independently of the piston position of the cylinder.
A turbocharger coupled to at least one cylinder, wherein the speed at which blowdown of the cylinder occurs directly drives the turbine;
An internal combustion engine according to claim 4,
The turbocharger further comprises an electric auxiliary motor for allowing the compressed gas supplied to the intake port to operate regardless of blowdown gas from the exhaust port and regardless of engine operating speed. An internal combustion engine comprising:
A single crankshaft having a plurality of pin journals and two opposing cylinders having an inner end adjacent to the crankshaft, each cylinder being reciprocally mounted therein and sandwiched therebetween for combustion Two cylinders having inner and outer pistons forming a chamber, two push rods each drivingly connecting each inner piston to a corresponding pin journal on the crankshaft, and each outer piston each having a crankshaft Comprising two pull rods drivingly connected to another corresponding pin journal above,
One of the two cylinders is arranged on the side where the intake port is close to the crankshaft, the exhaust port is located on the far side, and the other is arranged on the side where the intake port is far from the crankshaft, and the side where the exhaust port is close,
The internal combustion engine characterized in that the geometry and mass of these parts are selected to minimize the mechanical imbalance during operation of the internal combustion engine.
There are two pull rods for each cylinder, the two pull rods are on either side of the cylinder and are swingable to the inner end surrounding the associated crank pin of the crank shaft and the distal end of the associated outer piston, respectively. An internal combustion engine having an end remote from the coupled crankshaft.
The internal combustion engine according to claim 1, wherein the cylinder has a parallel longitudinal axis, and each axis is offset in a direction perpendicular to the axis of the crankshaft in a direction opposite to the axis .
Further, the turbocharger is connected to an exhaust port of the associated cylinder, undergoing blowdown gas from the exhaust port, it is connected to the intake port of the associated cylinder, for supplying compressed air to the intake port, wherein the this An internal combustion engine.
The turbocharger is connected to an intake impeller and includes an electric auxiliary motor for allowing the intake impeller to supply compressed air to the intake port regardless of blowdown gas. organ.
Each inner piston has a smooth end surface convexly curved in a plane perpendicular to the longitudinal axis of the crankshaft at the end remote from the combustion chamber, and an associated push rod assembly is connected to one of the crankshafts on the crankshaft. of which is connected to the crank pin, comprising a connecting rod having an outer end surface of the concave of the curved end surface and slidably engage the inner piston, the convex inner piston effective length of each push rod to the associated An internal combustion engine including a radius of a curved end surface.
Each inner piston has a smooth end surface curved convexly in a plane perpendicular to the longitudinal axis of the crankshaft at the end remote from the combustion chamber, and the associated push rod assembly is connected to one of the crankshafts. An internal combustion engine comprising a connecting rod coupled to a crankpin and having a concave outer end surface slidably engaged with the curved end surface of the inner piston.
A two-stroke internal combustion engine comprising opposing cylinders having opposing pistons,
1) A pair of opposing cylinders, each cylinder having two pistons reciprocally mounted therein, and the two pistons in each cylinder forming a combustion chamber therebetween,
2) a single crankshaft located in the middle between the two cylinders, the crankshaft having a plurality of crankpins ;
3) Each cylinder further
a) Each cylinder is adjacent to the single crankshaft and has an inner end and an outer end;
b) a cylinder wall having each cylinder intake port and exhaust port, wherein one of the pistons in each cylinder operates to open and close the exhaust port of the cylinder wall, and the other piston in each cylinder is , And operates to open and close the exhaust port of the cylinder wall, the intake port of one cylinder is located closer to the outer end of the cylinder, and the intake port of the other cylinder is closer to the outer end of the cylinder The exhaust port is located closer to the inner end of the cylinder,
c) the cylinder wall further has one or more slots from the outer end;
4) An electric auxiliary turbocharger connected to each cylinder, wherein the motor operates an intake air compressor to supply compressed intake air to the intake port,
5) a pair of push rod assemblies, wherein one push rod assembly transmits the pushing force from the innermost piston in each cylinder to the crank pin of the crank shaft;
6) A pair of pull rod assemblies, wherein one pull rod assembly transmits the pulling force from the outermost piston in each cylinder to different crank pins of the crankshaft , and the pull rod assembly is further connected to the slot in the cylinder wall. Communicate with the piston through
7) The Kurankupi down the internal combustion engine, wherein the balance of mechanical forces within the internal combustion engine is angularly positioned so as to substantially take.
A two-stroke internal combustion engine comprising opposing cylinders having opposing pistons according to claim 13,
The electric auxiliary turbocharger further includes:
An exhaust impeller coupled to the exhaust port of the cylinder, the blowdown operating the impeller, the impeller being connected to an intake compressor, and compressed air being supplied to the cylinder through the intake port A characteristic internal combustion engine.
The internal combustion engine further comprises means for directly injecting fuel into the combustion chamber formed between the two pistons of each cylinder.
Internal combustion engine, wherein the electrical auxiliary motor is an induction motor or switching 換Ri reluctance motor.
The internal combustion engine according to claim 9, wherein the turbocharger supplies compressed air at a maximum pressure immediately before the intake port is closed.
The internal combustion engine according to any one of claims 2, 3, 5, 10, and 16, wherein the internal combustion engine converts the energy of exhaust gas into electric energy by a turbocharger, and converts the energy into an electric storage device. An internal combustion engine configured to be rechargeable and configured to start the turbocharger using the stored electrical energy.
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All Through the Night is our November book and continues the war theme, but wasn't published until the war was nearly over in 1945.
When it arrived at the Freeport Public Library that year, the synopsis-writer was obviously not a GLH fan. "Cinderella's grandmother dies leaving her to counter her aunt and two mean cousins with the help of a soldier's love."
This book was one of Grace's last few stories. See if you agree or disagree with the 1945 reviewer. Has your vintage copy gone missing? Grab it in paperback. It's #6 in the Bantam or Tyndale Series.


Pac-12 teams will make up 25 percent of the Sweet 16 when the women's regional semifinals get started Friday. No. 2 seed Oregon State, No. 3 seed UCLA, No. 4 seed Stanford and No. 7 seed Washington all advanced through the first weekend of the tournament, giving the Pac-12 four teams in the final 16 for the first time in conference history. The league had never advanced more than three teams beyond the first weekend.
Oregon State romped into the regional semis a year after getting upset in the second round. UCLA and Stanford both held off upset bids on their home courts. And Washington pulled the surprise of the first two rounds by taking out two-time Final Four participant Maryland on the Terrapins' home floor.
The only stumble came from No. 2 seed Arizona State, which saw its run end with a second-round loss to perennial power Tennessee.
While the Pac-12 fell on its face in the men's tournament with six of seven teams eliminated on the opening weekend, the women's teams are living up to their billing. The SEC is the only other conference to advance four teams to the regional semifinals.
OREGON STATE: A season ago, the Beavers were the upstarts looking to the NCAAs as the next step in their progression to becoming an elite program, only to get upset in the second round on their home floor by Gonzaga.
That lesson has hung with Oregon State all season and when the same situation presented itself again, the Beavers made quick work of St. Bonaventure.
The Beavers have what appears to be the easiest matchup of the remaining Pac-12 teams, facing No. 6 seed DePaul in the regional semis, but a win there would likely set up a regional final against Baylor in Dallas.
UCLA: The Bruins are in the Sweet 16 for the first time in 17 years thanks to guard Jordin Canada. Held in check by South Florida for three quarters of the second-round matchup, Canada scored 17 of her 19 points in the fourth quarter as the Bruins held on.
The last time UCLA reached the regional semifinals was 1999, when the Bruins lost to Louisiana Tech in the regional final. The Bruins' reward for getting this far is facing No. 2 seed Texas in Bridgeport, Connecticut, and the winner likely getting a date with undefeated UConn.
STANFORD: The Cardinal needed a late rally and Lili Thompson's three-point play in the closing seconds to beat No. 12 seed South Dakota State in the second round. Stanford did not win a share of the regular-season conference title or the conference tournament for the first time since 2000. But the Cardinal are in the Sweet 16 for a ninth straight season.
Stanford has the most difficult challenge in the regional semis, facing No. 1 seed Notre Dame.
WASHINGTON: The Huskies are the surprise of the four after beating Maryland in the second round. It's been 15 years since the Huskies last made the regional semifinals, but with a scorer like Kelsey Plum anything is possible with the Huskies. Plum had 32 points in the win over Maryland and will need another big game when Washington faces No. 3 seed Kentucky on the Wildcats' home floor.
''We felt like the Pac-12 Conference schedule and conference tournament particularly had really gotten us prepared for (the Maryland) game and that weekend and the way it all unfolded it turned out to be 100 percent true,'' Washington coach Mike Neighbors said.