FIELD OF THE INVENTION

This invention pertains to a small displacement two-cycle internal combustion engine with a mechanical direct fuel injection system for powering portable power tools and equipment used in forestry, lawn, garden and construction, as well as small vehicles like scooters and mopeds.

BACKGROUND OF THE INVENTION

The advantages of two-cycle engines are well known. They are simple, have a high power/weight ratio, can be manufactured at a low cost and are very reliable. These characteristics have made the two-cycle engine the preferred power source for hand held appliances such as chain saws, line trimmers, leaf blowers and the like. However, the necessity for ensuring complete combustion and minimization of scavenging losses of the engine present significant problems.

Most of the modern two-cycle engines employ three basic types of scavenging systems: Loop scavenging (FIG

1

a

), cross scavenging and uniflow scavenging (FIG.

2

). The loop scavenging system, being the most popular due to its simplicity and effectivity. The scavenging ports direct a stream of air/fuel mixture into the cylinder, creating a loop like flow pattern, aiming to evacuate the remaining gases left from the combustion cycle. Despite the numerous improvements implemented through time in the loop scavenging system since its invention by Schnuerle in 1926, an unavoidable portion of unburned fuel is always released into the atmosphere as scavenging losses. This reduces the fuel efficiency of the engine and creates atmospheric pollution.

Pending and existing exhaust emission regulations imposed by the EPA and CARB on non-road equipment up to 19 Kw including lawn and garden equipment powered by internal combustion engines, strongly demand reduction in noxious substances such as hydrocarbons, nitrous oxides and carbon monoxide, in exhaust gas discharged mainly by two-cycle engines used on power tools and other lightweight applications.

In order to meet such existing and pending air pollution exhaust emission regulations, for such hand held two-cycle engines, much effort and expense has been directed in the last several years towards improving scavenging and fuel delivery systems for such engines to enable the same to meet such stricter exhaust pollution requirements, especially in regard to the unburned hydrocarbon (HC) component. In this field, the major hurdle has been to achieve this result at an affordable cost to the end user of such relatively low cost equipment, while also insuring that such engine improvements do not compromise the easy portability requirements for such engine powered handheld appliances and equipment.

New generations of lightweight four-cycle engines with low hydrocarbon emissions are among the technologies being developed for powering handheld portable tools. Their manufacturing cost, in-use emissions deterioration, serviceability and low power/weight ratio are still problems to resolve. An example of this technology is illustrated by U.S. Pat. No. 5,558,057.

Catalytic converters, fuel injection, uniflow scavenging and stratified scavenging are among the technologies aimed to reduce exhaust emissions in modern two-cycle engines.

Catalytic converters are well known from automotive applications as an efficient method to reduce exhaust emissions. The hydrocarbon reduction is a result of a chemical reaction that produces oxidation of exhaust gases. Unfortunately, the catalyst materials deteriorates with use, do not completely eliminates the hydrocarbon emissions and generates unwanted amounts of heat, factors that are not very appealing in small engines.

Uniflow scavenging is another method used to improve the fuel efficiency and to reduce the scavenging losses incurred in loop scavenged two-cycle engines. Uniflow engines were successfully used in the 30's on automotive and diesel engines by Trojan, Garelli, DKW, Puch, TWN and EMC. Longer scavenging loop and clever asymmetric geometry allowed the port timing of some split singles to be juggled so that the exhaust closes before charging has finished, which all helped to keep the fresh mixture out of the exhaust increasing the fuel efficiency. These advantages of uniflow scavenging are used for reducing exhaust emissions in modern applications. Examples of this method are provided by U.S. Pat. Nos. 4,079,705, 5,722,355 and 5,758,611

Another well-known approach successfully used to reduce scavenging losses is direct fuel injection systems. Thanks to recent electronic technology developments, electronic fuel injection is presently widespread as the preferred fuel delivery system in automotive applications. Unfortunately, this technology has not been commercially developed in low cost lightweight applications due to the electrical hardware required and its associated high cost. Also, the complexity of a fuel injection system to manage the small fuel volumes required at idle and at full throttle conditions, has remained as another serious obstacle to successfully implement direct fuel injection systems in hand held appliances powered by two-cycle engines.

Long before electronic fuel injection was technologically possible, mechanical fuel injection systems were widely used on diesel engines and on high performance gasoline engines. Many attempts have been made to adapt mechanical fuel injection to small engines, but cost and functional factors have been significant barriers to these systems.

During last century, the earliest efforts with regard to the development of a mechanical fuel injected two-cycle engine, was the Clerk engine. This engine used a mechanical pump to transfer air/fuel mixture to a working cylinder. Since then, other engine inventors followed the same basic Clerk's principles in their engines. Most of these early inventions were originated from diesel engines concepts where high compression ratios are necessary. The U.S. Pat. No. 607,276 by the Joseph Reid Gas Company issued in 1898 describes an engine with a pump cylinder and a power cylinder used for oil well service, where the pump cylinder was used to transfer natural gas mixed with air into the working cylinder. Another early application of volumetric fuel pumps in two-cycle engines were the supercharged racing engines by DKW and Schilha in the early 1900's.

The U.S. Pat. No. 1,168,425 by Rosenhagen issued in 1916 describes a typical example of prior art engines using a volumetric pump to transfer the air/fuel mixture into the working cylinder. This engine uses timing differences based on the radial positioning of a pump piston in relation to a power piston to create anticipated upward motion of the pump piston thus creating a pressure differential between both cylinders. Complicated valving and fuel passages, low speed, large air/fuel paths, high pumping losses and lack of lubricating means prevented the success of this invention seeking improved volumetric efficiencies.

Another examples of use of volumetric pumps to transfer a combustible mass to a working cylinder is described on patents by Houyez (France 908,916), Silander (Belgium 515,577), Kerrebrok (U.S. Pat. No. 4,506,634), Voisin (France 1,084,655) and Lepore (Italy 434,901) among others.

The aforementioned prior art on two-cycle engines with volumetric pumping systems was intended for low speed, large engines, capable of absorbing large pumping losses, high levels of vibration and with a multiplicity of components not tolerated by small hand held engines with low levels of power and inexpensive manufacturing. These prior art engines did not succeed due to the competitiveness of loop scavenged two-cycle engines in an era where exhaust emissions were unimportant. Therefore, there is a need in the art for a small high performance two-cycle engine with very low hydrocarbon emissions that can be successfully fabricated with current mass production methods at a cost affordable for such inexpensive applications.

A modern example of air assisted mechanical fuel injection systems using volumetric fuel pumps is provided by the FAST system. A volumetric pump driven by a secondary crankshaft introduces a rich air/fuel mixture into a power cylinder. The Italian manufacturer Piaggio successfully uses this system to reduce exhaust emissions in motor scooters. As it will be learned further in the description of this invention, the engine object of the present invention uses equivalent physical principles to those used by the FAST system to gain volumetric efficiency and reduced hydrocarbon emissions. The engine object of the present invention provides the same effects, but thanks to the use of a greatly simplified mechanical system, it allows a lightweight and compact construction as well as reduced mechanical losses that allows its utilization on portable tools.

It is obvious to the person skilled in the art, that the prior art of air assisted mechanical fuel injection in two-cycle engines, often has a complex and bulky construction not desirable for lightweight applications and hand held portable tools where compactness, low weight, low cost and low emissions are the dominating factors.

SUMMARY OF THE INVENTION

The object of the present invention is to provide a lightweight, compact and economical two-cycle engine that offers a power/weight ratio similar to conventional prior art loop-scavenged two-cycle engines but with very low scavenging losses, thus, with very low exhaust emissions. Its simplicity and purposeful construction substantially reduce all the problems found in prior art two-stroke air-assisted fuel injection engines allowing at the same time a low manufacturing cost as required in hand held gas powered tools as those used in construction, forestry, lawn and garden applications.

A two-cycle, crankcase scavenged internal combustion engine is provided comprising a siamese cylinder in which two pistons reciprocate parallel to each other. One of the cylinder bores contains the exhaust ports, intake ports, scavenging ports and a combustion chamber as in a typical single cylinder two-cycle engine. The combustion chamber is provided with a spark plug and an inlet valve. This cylinder bore in cooperation with a piston provides the power cycle. The second cylinder bore in cooperation with a slave piston provides the pumping action necessary to introduce a rich air/fuel mixture into the power cylinder. This second cylinder bore has at its top end an inlet valve and a passage communicating to the power cylinder. This communicating passage is positioned in order to obtain the smallest possible dead volume. The two pistons are connected to a common crankshaft by single or individuals connecting rods. This geometry provides an asymmetric motion of the pistons, which enables the timing of the pump piston to be considerable advanced in relation to the power piston. This timing advance produces a significant pressure differential between both cylinders, which is utilized to pump a rich air/fuel mixture into the power cylinder.

Another advantage of the asymmetric location of the crankshaft in relation with the centerline of the power cylinder is the reduction of mechanical friction. In typical two-cycle engines, during the power stroke the connecting rod angle in relation to the cylinder centerline determines the amount of the force applied by the piston skirt over the cylinder surface. This reaction is called the major thrust force. Significant friction force is generated over the major thrust surface and it is directly proportional to the angle between the connecting rod and the cylinder centerline. This frictional force produce heat and wear. In the engine object of the present invention the rod angle is maintained to small values during the expansion cycle. This allows maximizing the piston force transmitted to the crankshaft, while reducing the heat and wear generated by the friction .

With the arrangement described above, the siamese cylinder two-cycle engine operates as a typical loop scavenged engine within the power cylinder, but with significant reduction of pollutants into the exhaust gases. This significant reduction of pollutants into the exhaust gases is accomplished by the combined action of several improvements. As aforementioned, in typical two-cycle engines, during the scavenging period a portion of scavenging gases always escape through the exhaust port; as the engine object of the present invention is also a loop scavenged two-cycle engine within the power cylinder, it will have scavenging losses. But due to the use of pure air to scavenge the exhaust gases, some of this air escaping through the exhaust port as scavenging losses, mixes with the exhaust gases, which contains high levels of carbon monoxide. As a result of this chemical reaction, the excess air into the exhaust gases stream oxidizes significant amounts of carbon monoxide. The carbon monoxide is then transformed into carbon dioxide, which is a harmless gas.

As the air/fuel mixture is injected directly into the combustion chamber of the power piston after the exhaust port has been closed, virtually there are no raw fuel losses into the exhaust gases stream, therefore hydrocarbon emissions are almost eliminated. Another advantage not obvious with the unfamiliar with the combustion process, is the reduction of fuel droplet size caused by fuel atomization into the carburetor venturi and subsequent expansion through the injection valve into the combustion chamber of the power cylinder. The final expansion process causes a droplet size much smaller than in current carbureted systems or direct fuel injection systems, therefore an improved and complete combustion process is enabled due to the increased interaction between fuel and air molecules. Another advantage yet, is the resulting stratified volume of rich air/fuel mixture injected around the spark plug at the end of the compression cycle. This stratification is known to promote combustion initiation and flame propagation.

As the engine object of the present invention is a combination of a traditional two-cycle engine with an integrated mechanical fuel injection system, the configuration of the two-cycle engine may be configured as other two-cycle engine including reed valve systems, piston ported systems, rotary valve systems, and combination thereof without departing from the spirit of the present invention.

The preferred embodiments of this invention have several inventive aspects, which jointly contribute to the main functional object of the invention: to reduce exhaust emissions and improve fuel efficiency while preserving the desired features of a typical two-cycle engine. One of the embodiments describe the use of a low cost and compact air assisted mechanical fuel injection system adapted to a small two-cycle engine to power a portable tool. Another embodiment describes the use of the asymmetrical timing well known in uniflow engines, typically used at the bottom dead center to create a pre-outlet of the exhaust port. Instead, it is used in the present invention at the top dead center to create a pressure differential for allowing fuel injection. Another further embodiment shows how the general construction and design of the engine allows simple manufacturing and low parts count, not typical for prior art air assisted mechanical fuel injected engines.

The invention will be better understood and further objects and advantages thereof will become more apparent from the ensuing detailed description of preferred embodiments taken in conjunction with the drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention will be described in closer detail in the following by means of various embodiments thereof with reference to the accompanying drawings, wherein identical numeral references have been used in the various drawing figures to indicate identical parts.

FIG. 1

shows a schematic illustration of prior art two-cycle engines.

FIG. 2

is a side view of a power tool having a low emission two-cycle engine embodying the present invention.

FIG. 3

is a cross sectional view of the engine embodying the present invention. The sectional view is through the engine's cylinder centerline.

FIG. 4

is cross sectional view of the engine of FIG.

3

. This sectional view is through the crankshaft axis and the cylinder axis.

FIG. 5

shows several connecting rod configurations adaptable to the engine object of this invention.

FIG. 6

illustrates the different operating stages of the engine object of this invention.

FIG. 7

shows port timing diagrams for conventional two-cycle engine and for the engine object of the present invention.

FIG. 8

shows a pressure vs. crank angle diagram during one crank revolution of the engine object of this invention.

FIG. 9

illustrates the constructive details of a one-piece cylinder head and a cylinder with detachable head.

FIG. 10

shows a cross sectional view of the transfer valve unit.

FIG. 11

shows an engine with the crankcase induction and pump induction controlled by piston displacement.

FIG. 12

shows a dual level cross section perpendicular to the cylinder axis and through the intake and exhaust ports of an engine with piston ported intake system.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS.

The embodiments of the present invention will now be explained with reference to the accompanying drawings.

Referring to

FIGS. 2

,

3

and

4

, the invention is described in connection with a gas-powered line trimmer. The gas-powered trimmer is intended to be representative of a hand held portable tool where the engine object of this invention may be used as the power source. Some of the commonly known portable tools where this low emission two-cycle engine may be utilized are chain saws, blowers, cultivators, edgers, hedge trimmers, snow blowers, and the like. It also may be understood that the use of this engine is not limited to any other applications where the use of a conventional two-cycle engine is advantageous.

In a known manner, the low emission two-cycle engine object of this invention is used to drive the work implement of the gas powered line trimmer shown by

FIG. 2

, which is represented by a cutting head

12

. The rotational power is transmitted to the cutting head

12

by a flexible shaft

13

disposed into a rigid tube

14

. The engine housing

15

is secured to the tube by a clamp

67

into the nose portion of an engine housing

15

. At the end of the pocket where the tube is nested inside the nose of the engine housing

15

, is located a clutch drum element

16

including a coupling section

17

to receive the end of a drive shaft

13

. Disposed into the clutch drum

16

is a clutch shoe assembly

18

and a clutch hub

19

. The clutch hub

19

and the clutch shoe assembly

18

are mounted on a crankshaft

34

. At low engine speeds the clutch drum

16

is disengaged from the cutting head

12

. When the engine reaches certain rotational speed, the clutch shoe

18

engages the drum

16

and couples the engine to the work implement represented by the cutting head

12

. The gas-powered trimmer has two handles

20

and

21

for a person to hold and maneuver the trimmer to cut vegetation. Near the main handle

20

is located a throttle control

22

for a person to control the cutting speed of the work implement.

FIG. 3

shows a cross section through the cylinder bores and

FIG. 4

shows a longitudinal section through the crankshaft of the low emission two-cycle engine object of this invention comprising: a cylinder block

37

including Siamese cylinders and a combustion chamber

46

. A first cylinder called hereunto the power cylinder

45

and a second cylinder called hereunto the pump cylinder

61

. Both cylinder bores are parallel to each other and spaced by a common wall

43

. The plane, which contains the centerline of the cylinder bores, is perpendicular to the crankshaft axis. The cylinder block

37

is preferably cast as a single piece of aluminum alloy.

Cooling fins

44

for dissipating the heat generated by the engine surround the cylinder block

37

. Disposed into the cylinders are two pistons: a power piston

40

and a slave piston

59

, which are axially guided into the power cylinder

45

and into the pump cylinder

61

respectively. The power cylinder

45

in cooperation with the power piston

40

provides the power cycle. The power cylinder

45

comprises the exhaust port

41

, the scavenging ports

39

and a combustion chamber

46

. It will be further described that the power cylinder

45

may also comprise at least a crankcase intake port laterally opened into the cylinder wall when the engine is configured with a piston ported induction system. At the top portion of power cylinder

45

is provided a combustion chamber

46

defined by the end face of the power piston

40

dome and the end surface of the power cylinder

45

. Provided inside the combustion chamber

46

is a spark plug

47

face and a transfer valve

48

face. The spark plug

47

is mounted to the cylinder block by threaded means and extends into the combustion chamber

46

. The transfer valve

48

is spring loaded by means of a compression spring

79

(FIG.

10

). When the cylinder block

37

is manufactured in one piece, due to the small bore size, it may be difficult to machine and place the transfer valve from the inside of the cylinder. For such purposes, a transfer valve bore

49

is drilled from the outside of the cylinder extending into the combustion chamber

46

. This allows the precise placement of a transfer valve unit

77

shown in FIG.

10

. This transfer valve unit is an integral part consisting of a transfer valve body

78

, a poppet type transfer valve

48

, one compression spring

79

, spring retaining means

80

and a cap

81

. The compression spring

79

maintains the transfer valve

48

closed and it is calibrated to open when the pressure differential between both cylinders reaches predetermined levels. The cap

81

seals the open end of the transfer valve body

78

and prevents any fluids from escaping to the atmosphere. The transfer valve unit

77

has a lateral opening

82

that communicates the power cylinder

45

with a transfer passage

50

. The transfer passage

50

communicates the power cylinder

45

with the pump cylinder

61

. This transfer passage

50

, allows the load of rich art/fuel mixture compressed into the pump cylinder

61

to be transferred into the combustion chamber

46

of the power cylinder

45

after biasing the spring force acting over the transfer valve

48

.

The power piston

40

design and construction is identical to those used in conventional two-cycle engines.

As shown by FIG.

9

and

FIG. 11

, to simplify manufacturing, the transfer passage

50

is purposely located to be drilled from the outside of the cylinder block

37

through the inlet valve block mounting opening

69

at the top of the pump cylinder

61

or through the spark plug opening if a piston ported configuration is used. Another configuration that minimizes manufacturing cost is to locate the transfer valve bore

49

intersecting the nearest corner of the pump cylinder

61

. It is very important that the transfer passage

50

is designed with the minimum possible length and volume in order to reduce the dead spaces during compression, which reduces pumping losses.

The pump cylinder

61

in cooperation with the slave piston

59

, works as a volumetric pump to transfer a rich air/fuel mixture into the power cylinder

45

. Due to the small amounts of rich air/fuel mixture needed for the combustion process into the power cylinder

45

, the diameter of the pump cylinder

61

is substantially smaller in size than the power cylinder

45

. The diameter of the pump cylinder

61

is inversely proportional to the degree of concentration of fuel into the air and must be minimized to reduce pumping losses. The pump cylinder

61

contains a pump chamber

68

defined by the surface of the dome of the slave piston

59

, the end surface of the pump cylinder

61

and the face of a inlet valve housing

53

. The inlet valve housing

53

includes a reed type inlet valve

51

attached to the surface of the inlet valve housing

53

facing the pump cylinder inlet valve opening

69

. The inlet valve housing

53

is disposed over the top portion of the pump cylinder

61

. The inlet valve housing

53

comprises an inlet passage

52

for fluid communication of a carburetor

56

with the inlet valve

51

. The carburetor

56

is attached to the inlet valve housing

53

at the end opposite to the inlet valve

51

. The material of the inlet valve housing

53

has a low coefficient of heat transmission to avoid engine heat migration towards the carburetor

56

. For the shown configuration, the angular position of the inlet valve housing

53

in relation to the cylinder axis, allows the convenient location of the carburetor

56

and an air filter box

58

. The angular positioning of the inlet valve housing allows also an updraft flow of the air/fuel mixture, avoiding engine flooding and fuel puddles during starting. The fuel inlet valve

51

is a reed type valve with elastic properties urging one of the faces of the valve to lay against the opening of the inlet valve housing

53

. When the slave piston

59

starts its descending stroke within the pump cylinder

61

, a negative pressure is created, allowing air/fuel at a higher pressure to bias the elastic force acting on the reed inlet valve

51

, opening the air/fuel mixture flow into the pump cylinder

61

.

As described in

FIG. 11

, the reed valve inlet system may be configured as a piston ported system where the air/fuel inlet port is located in the side wall of the pump cylinder

61

, wherein opening and closing of the air/fuel flow is controlled by the displacement of the upper edge of the slave piston dome. Furthermore, the crankcase inlet port

75

can be also located in the wall of the pump cylinder, wherein opening and closing of the port is controlled by the position of the lower edge of the slave piston

59

skirt. It will be also shown that the crankcase inlet port

75

, can be located in the side wall of the power cylinder

45

, wherein opening and closing of the port, is controlled by the displacement of the lower edge of the power piston skirt, as in traditional two-cycle engines.

The slave piston

59

has a dome surface that follows closely the shape of the top of the pump cylinder

61

and the inlet valve

51

face. As aforementioned, the reduction of the dead spaces during compression is necessary to reduce compression losses, therefore, increasing engine efficiency. A ring is disposed in the cylindrical portion of the slave piston

59

to prevent the leakage of gases through the clearance space between piston and cylinder. As the slave piston

59

is intended to withstand only compression cycle forces, lightweight materials like aluminum, magnesium or carbon matrix composites can be used to further reduce the weight of the reciprocating masses, which is important to reduce engine vibration. As minimization of vibration and reduction of engine package is necessary, a very short connecting rod is required and cuts into the cylinder walls are necessary to allow space for the connecting rod motion.

The cylinder block

37

is disposed over an engine block

32

containing a crankcase chamber

31

and the main bearings bore. The engine block

32

preferably cast in one-piece of a lightweight material like magnesium or aluminum. The engine block walls and the rear crankcase plug

66

comprise the boundaries of the crankcase chamber

31

. The engine block

32

contains the main bearings

70

and

73

on which the crankshaft

34

is rotatively mounted. Attached to the main bearing bore and in between the main bearings

70

and

73

is located a crankcase seal

72

which has the inner lip in contact with the crankshaft

34

to prevent air/fuel leaks. The crankshaft

34

has a counterweight-crank

33

disposed at the end contained into the crankcase

31

. A crankpin

35

is eccentrically disposed into the counterweight-crank

33

, being parallel to the crankshaft

34

. Angularly opposed to the crankpin location is the heavier mass of the counterweight-crank

33

, which is used to counterbalance the reciprocating mass of the engine. A connecting rod

36

is rotatively mounted over the crankpin

35

by means of a crankpin bearing

30

. One arm of the forked connecting rod

36

called hereunto the main arm

38

, connects with the power piston

40

, the other arm called hereunto the slave arm

60

, connects with the slave piston

59

. The connecting rod arms are connected to the slave and power pistons by the wristpins

54

and

55

respectively. Wristpin bearings

29

are sandwiching between the wristpin and the wrispin head eye of each of the connecting rod arms to reduce friction. The wristpins

54

and

55

are parallel to each other and to the crankpin

35

.

At the opposite end of the crankshaft

34

containing the counterweight-crank

33

, is disposed a flywheel-fan

71

, which has three major functions: First, to provide the necessary inertial forces required during the compression cycle. Second, to provide the necessary airflow required for cooling the cylinder block

37

. Third, to provide in cooperation with the ignition module the electrical current required for spark generation. The skilled in the art will recognize that many different combinations of crankshaft systems as double supported, double counterweight crankshaft, double supported single counterweight crankshaft, can be utilized without departing from the spirit of the invention.

The slave piston

59

is pivotally connected to the crankshaft

34

through the slave arm

60

of the connecting rod

36

. The upper portion of the slave arm

60

is connected to the slave piston

59

by means of the wristpin

54

. A wristpin bearing

29

is sandwiched between the wristpin-head eye and the wristpin

54

. The lower end of the slave arm

60

is attached to the crankpin-head of the connecting rod

36

. The neutral axis of this slave arm

60

is coplanar with the plane formed by the crankpin

35

axis and the wristpin

54

for maximum column strength as shown in

FIG. 5

a

. Wristpin bearings may be replaced by the bearing surface of the connecting rod arm eye if the specific loads are low.

FIG. 5

a

shows the connecting rod

36

as a one-piece element comprised by a slave arm

60

, the main arm

38

, the slave arm wristpin head

84

, the slave arm wristpin eye

85

, the main arm wristpin-head

86

, the main arm wristpin-eye

87

, the crankpin-head

88

and the crankpin-eye

89

. The main arm

38

is attached to the power piston

40

and the slave arm

60

is attached to the slave piston

59

. The neutral axis of both arms must be in the same plane of the crankpin

35

axis. This is very important to avoid twisting forces around the crankpin during the expansion cycle. At this point, when the main arm

38

is under high compressive forces produced by the combustion gases expansion, the slave arm

60

is under tensile forces due to the suction applied by the slave piston

59

into the pump cylinder

61

.

A typical problem found in twin parallel cylinder uniflow engines is the variation of the wristpin centers when the angular position of the connecting rod changes during the crankshaft rotation. Methods to compensate for wristpin distance variation were used on early uniflow engines during the 1930's. These engines used a forked connecting rod in which the arms were allowed some flexibility (Trojan) or mother-slave rod configurations with intermediate linkage (Zoller). These engines also had very long connecting rods into a large crankcase with very low pumping efficiency. In the engine object of the present invention, like in uniflow engines, the variation of wristpin centers is a problem to overcome. By allowing flexion of one of the connecting rod arms the same results are achieved. The connecting rod arms are designed as short as possible to minimize crankcase volume to maximize the pumping efficiency but yet must be flexible enough to accommodate the variation of the wristpin centers.

As the power piston

40

is exposed to the force of the combustion gases, the main arm

38

is designed for high column strength to withstand such compressive stresses similar to conventional two-cycle engines. On the other hand, the slave piston

59

is substantially smaller than the power piston

40

, it is only exposed to compression cycle gas pressure, therefore, the magnitude of the force transmitted to the slave arm beam

60

is relatively low, allowing to design the slave arm beam with a fairly thin section. This thin section allows good levels of flexibility without reaching the fatigue limits for the material. The flexion of the slave arm allows accommodating the wristpin distance variations, while the main arm is substantially rigid.

FIG. 5

a

also shows the cross sectional view of both connecting rod arms. This view illustrates how the moment of inertia in the flexing plane of the slave arm is much smaller than the moment of inertia of the main arm. This orientation allows the minimization of the flexural stresses while allowing good column strength.

In its natural state, the position of the slave arm wristpin-eye

85

center in relation with the main arm wristpin-eye

87

center, is in an intermediate position between the maximum and minimum distance variation between wristpin centers. For instance if the maximum wristpin distance is 35 mm and the minimum distance is 34 mm., the natural state distance between both wristpin eye centers must be around 34.5 mm. This allows minimum bilateral flexion of the slave arm during operation. By allowing minimum flexural stresses, the reaction force over the wristpins is also reduced, therefore, the friction force between piston and cylinder is also reduced.

The connecting rod is preferably made of a lightweight metal in order to reduce the engine vibration. As previously mentioned, the length of the connecting rod is maintained to a minimum in order to reduce the dead spaces into the crankcase, necessary to increase the scavenging gases pressure and to avoid crankcase pumping losses. Alternate connecting rod designs are possible without deviating of the main scope and spirit of the present invention as shown by

FIG. 5

b

and

5

c.

Another embodiment of the present invention is a method to compensate for the wristpin distance as shown in

FIG. 5

b

. An eccentric collar [

100

]

83

is sandwiched between slave piston wristpin

54

and the slave arm wristpin-head eye

85

. Due to the eccentricity of the wristpin mounting location in relation to the center of the wrispin-head eye center, little change in distance in the plane parallel to both wristpins are compensated by rotation of the eccentric collar

83

within the wristpin-head eye

85

. The eccentric collar

83

is located in such a way that its outside diameter is in contact with the internal diameter of the wristpin-head eye

85

. Its internal diameter which is eccentrically located in relation with its outside diameter, is in contact with the wristpin

36

. Proper lubrication is required to reduce friction forces when the eccentric collar

83

is in direct contact with the cooperating surfaces. Needle bearings may be used over the surfaces in contact to further minimize friction between sliding surfaces without deviating of the main scope of the invention.

FIG. 5

c

shows two individual connecting rods achieving the same function as the one piece forked connecting rod

36

. Both connecting rods

91

and

92

are mounted over the same crankpin

35

. The centerline of the power cylinder

45

and the pump cylinder

61

is offset in such a way that the plane of motion of the power piston

40

and the slave piston

59

are in the same plane of the centerline of its corresponding rod. It is also possible to connect both rods in a mother-slave rod configuration as used in multi-cylinder radial engines, but this will add prohibitive vibration, weight and cost to the engine.

The cylinder block

37

and the engine block

32

has been described as different elements, but it is always possible combine them as a one-piece casting in configurations where manufacturing cost is the main concern.

With the aforementioned engine elements, when the combustion process is initiated into the power cylinder

45

, the rapidly expanding gases move the power piston towards the bottom dead center position. The rectilinear motion of the power piston

40

, is transmitted to the crankpin

35

by the connecting rod main arm

38

, then converted into the rotary motion by the crankshaft

34

. Utilizing the energy created by the power piston

40

, the slave piston

59

connected to the same crankpin

35

by the connecting rod slave arm

60

, moves downwards creating a negative pressure within the pump chamber

68

. As the negative pressure builds up, the air/fuel mixture crosses through the fuel inlet valve

51

, by biasing the spring force that keeps it closed against the face of the inlet valve housing

53

. This suction stage is illustrated by

FIG. 6

a

. The suction stage within the pump cylinder

61

, is simultaneous with the power stroke within the power cylinder

45

. Also should be noted the small rod angle within the power piston

40

, which enables most of the piston force to be transmitted to the crankpin without high trust forces over the piston and cylinder walls as illustrated by FIG.

11

.

The fuel inlet valve block

53

comprises a fuel inlet passage

52

for fluid communication with the carburetor

56

. The carburetor

56

contains fuel flow metering means

57

which is synchronized with the crankcase intake throttle valve

63

which regulates the flow of the scavenging fluid into the crankcase chamber

31

. As the slave piston

59

reaches its bottom dead center position, the suction force decreases, the air/fuel mixture flowing through the fuel inlet valve

51

stops and the valve

51

closes. At this stage the pump cylinder

59

is completely filled with a rich air/fuel mixture. At the same time within the power cylinder

45

the power piston

40

is in the proximity of its bottom dead center position, opening the exhaust port

41

. This allows the combustion gases pressurized into the power cylinder

40

to be released into a muffler

42

for their further releasing into the atmosphere. Immediately afterwards, the air into the crankcase chamber

31

, already compressed by the descend of the power piston

40

and the slave piston

59

, is released into the power cylinder

40

through the scavenging ports

39

located in the side of the wall of the power cylinder

45

. The scavenging gas then completes the evacuation of the residual exhaust gases left into the power cylinder

45

. This is commonly known as the scavenging cycle in conventional crankcase scavenged two-cycle engines and it is illustrated by

FIG. 6

b.

As typically found in loop scavenged two-cycle engines, some of the scavenging gases escape through the exhaust port

41

allowing raw hydrocarbons to be released into the atmosphere, creating an air pollution problem. There are also scavenging losses in the engine object of the present invention. Unlike to the effect of the scavenging losses in conventional two-cycle engines, the scavenging losses in this air scavenged engine, are beneficial. The oxygen contained into the scavenging air, mixes with the exhaust gases allowing the reduction of carbon monoxide into carbon dioxide, which is a harmless gas.

As shown by

FIG. 6

c

, after the scavenging cycle is completed, the compression cycle begins. When the slave piston

59

and the power piston

40

move upwardly driven by the inertial forces of the rotating masses of the engine, the power piston

40

starts compressing the remaining of the scavenging gas trapped into the power cylinder

45

. Simultaneously, the slave piston

59

starts the compression of the rich air/fuel mixture admitted during the previous cycle. The pressure differential between the two cylinders starts building up rapidly due to the asymmetrical configuration which allows the slave piston

59

to anticipate its upward motion within the pump cylinder

61

relative to the power piston

40

. FIG.

8

.

FIG. 7

shows a schematic illustration of the port timing in a typical two-cycle engine and the port timing of the engine object of the present invention.

FIG. 7

a

, shows the port timing diagram for a conventional two-cycle engine. It must be noted that in a conventional two-cycle engine, the opening and closing of the ports occurs at the same angle of crank rotation before and after the top dead center (TDC) position and bottom dead center (BDC) position. This is due to the positioning of the crankshaft directly under the centerline of the cylinder, which allows the piston to travel the same distance at same crank angles from the TDC-BDC line. This special configuration is called symmetrical port timing and as shown by

FIG. 7

a

. The shaded areas representing the exhaust and scavenging ports opening period, are symmetrical in both sides of the TDC-BDC line. When the crankshaft is moved away from the plane of the cylinder centerline, the port timing is shifted to one side of the TDC-BDC line. Under this condition, the opening and closing of the ports occurs at different crank angles. This geometry is called asymmetrical timing.

FIG. 7

b

shows how the shaded areas representing the exhaust and scavenging ports opening period, are shifted to the right side of the [line] TDC-BDC line. Then the pistons reach BDC and TDC at different crank angle. This crank angular difference is called the phase shift “Z”. The intake period within the power cylinder, is replaced by the injection period (

FIG. 7

c

), which occurs after the exhaust ports are closed and before the crank reaches TDC position. The magnitude of the phase shift angle “z” is very important for the proper function of the engine object of the present invention. Systems with phase shift angle under 15 degrees will not operate with the performance factors required to meet the applications requirements.

The fuel intake period within the pump cylinder has two phases: Induction and compression phase. During the induction phase the descending motion of the slave piston within the pump cylinder creates a negative pressure differential that allows air/fuel mixture from the carburetor to enter into the pump cylinder. During the ascending motion of the slave piston, it compresses the previously admitted fluids, forcing them into the power cylinder through the transfer valve

48

. This is called the compression phase.

FIG. 7

c.

Asymmetrical timing is commonly used in typical split uniflow engines, as shown in

FIG. 1

b

. Uniflow engines take advantage of this kinematics to obtain a considerable advance of the opening of the exhaust port over the scavenging ports. The same geometry produces a phase shift at the TDC position which the leading piston to reach its TDC first than the trailing piston. This phase shift at TDC is of no benefit to split uniflows. On the other hand, the engine object of this invention, uses the anticipation of the leading piston at TDC to create a substantial pressure differential. This pressure diferential allows to transfer air/fuel mixture from the pump cylinder towards the power cylinder. This engine also take advantage of the BDC phase shift to improve the air trapping efficiency.

The volumetric compression ratio of the power cylinder-power piston combination is substantially smaller than the volumetric compression ratio of the pump cylinder-slave piston combination. This allows greater gas pressure within the pump cylinder than in the power cylinder during the compression cycle. The combustion chamber

46

volume is calculated to receive the swept volume of the power piston

40

, added to the swept volume of the slave piston

59

, in such a manner that the final gas compression values do not exceed the typical compression ratio of conventional two-cycle engines. It may be noted by the skilled in the art the obvious supercharging abilities of this engine. In utility engines, as gasoline is the preferred fuel, the final compression ratio into the power piston must no exceed certain values that may allow pre-ignition.

When the slave piston

59

approaches its top dead center, the pressure of the air/fuel mixture reaches levels substantially higher than the gases into the power cylinder. This pressure differential allows biasing the spring force holding the transfer valve

48

closed. This allows the rich air/fuel mixture to enter into the combustion chamber

46

. Once the slave piston

59

reaches its top death center position, the air/fuel flow stops and the transfer valve

48

returns to the closed position. The fluid transferred into the combustion chamber

46

produces a stratified load of rich air/fuel around the spark plug, which is known to improve the initiation of the combustion process and improve flame propagation.

Immediately after the transfer valve

48

closes, the spark ignition is initiated within the combustion chamber

46

. This occurs when the power piston is between 30 and 15 degrees before its top dead center as in conventional two-cycle engines. The rapid expansion of gases into the power cylinder caused by the combustion process produces the displacement of the power piston initiating the power stroke all over again as illustrated in

FIG. 6

e.

For better understanding of the gas dynamics occurring during the engine operation,

FIG. 8

illustrates the change of gas pressure into the cylinders during one revolution of the engine. It can be observed how the pressure levels within the pump cylinder are substantially greater than into the power cylinder during the compression cycle. The different pressure values within the power cylinder and the pump cylinder during the compression phase are due to the pressure resistance to open the transfer valve

48

. It can also be noted the effect of the combustion process causing the higher-pressure values within the power cylinder immediately after the slave piston reaches its TDC and the combustion is initiated.

As illustrated by

FIG. 9

a

, the cylinder block

37

containing the two cylinder bores

45

and

61

is designed to allow high volume manufacturing processes such as die-casting. The head has been built as an integral part of the cylinder block

37

to reduce weight and cost. All the features at the top of both cylinder bores can be specifically designed to allow die cast and simple machining operations for a low cost end product. Following this manufacturing criteria, the transfer valve unit

77

can be built as a separate assembly which is fitted into the cylinder block after machining of the transfer valve bore

49

.

An alternate method of construction of the cylinder head without deviating from the spirit of the present invention, is using a detachable head as shown in

FIG. 9

b

in which two parts replace the cylinder block

37

: the cylinder bore block

98

and the head block

99

. In this configuration, there is more accessibility to the surfaces to be machined. As a result of this combination, the transfer valve elements and the transfer passage can be easily machined into the head block

99

. Also the inlet reed valve

51

and the inlet passage

52

are easily placed on the cast structure.

As aforementioned, the engine object of the present invention is basically a two-cycle engine with the cantilever crankshaft system. Double supported crankshaft systems may be utilized without departing from the spirit of the present invention. Also, there is not limitation to utilize any of the scavenging or induction systems typically used by these engines such as crankcase scavenging, external scavenging, piston ported induction, reed valve induction or rotary valve induction. Following, some of these possible configurations are described.

FIG. 3

shows an engine configured with reed induction system, where the scavenging gas flow is controlled by the crankcase intake reed valve

65

mounted over the reed block

64

. A crankcase throttle valve

63

controls the rate of flow of the scavenging gases entering the crankcase. The crankcase throttle valve

63

is mechanically linked to the throttle valve

57

of the carburetor

56

. Both air streams entering the engine through the carburetor

56

and the reed block

64

are connected to an air filter box

58

. An air filter

62

is disposed into the air filter box

58

. The same functionality can be also obtained by using a double-barrel carburetor, where the crankcase intake throttle valve

63

is disposed into the secondary barrel in fluid communication with the reed block

64

.

FIG. 11

shows an engine where the air/fuel intake reed valve

51

and the crankcase intake reed valve

65

, have been replaced by a piston ported valve system. The air/fuel intake port

52

and the crankcase intake port

75

are in fluid communication with a double barrel carburetor

95

, where the primary barrel

97

contains the means for providing a rich air/fuel mixture. Adjacent to the primary barrel

97

is the secondary barrel

96

including the crankcase intake throttle valve

63

. When the slave piston

59

starts its descending motion towards its bottom dead center position, a vacuum into the pump chamber

68

is created. As soon as the top edge of the piston dome uncovers the air/fuel intake port in the side wall of the pump cylinder, a rich air/fuel mixture rushes into the pump chamber

68

initiating the pump induction_cycle. When the slave piston starts its ascending motion towards its top dead center position, it covers the inlet valve

52

stopping the air/fuel flow into the pump chamber

68

, initiating the pump compression cycle. When the slave piston is near its top dead center position, the lower edge of the slave piston skirt uncovers the crankcase intake port

75

, allowing air to enter into the crankcase

31

. When the slave piston starts its descending motion towards the bottom dead center position, it covers the crankcase intake port

75

, initiating the crankcase compression cycle.

FIG. 12

shows another version of the engine object of the present invention, also configured with piston ported inlet valves. Similar to conventional two-cycle engines, the crankcase intake port

75

is disposed in the side wall of the power cylinder

45

in fluid communication with the crankcase

31

, wherein the displacement of the lower edge of the skirt of the power piston

40

controls the opening and closing of the intake port

75

. The scavenging ports

39

are rotated to allow the placement of the crankcase intake port

75

through the wall of the power cylinder

45

. The exhaust port is centered between the two scavenging ports

39

. This configuration also allows the use of a piston ported air/fuel induction system similar to the system shown in FIG.

11

. Similarly, the air/fuel intake is in fluid communication with the primary barrel

97

of a double barrel carburetor

95

containing means to supply a rich air/fuel mixture. Also, like in the engine of

FIG. 11

, the crankcase intake port is connected to the secondary barrel

96

of the double barrel carburetor

95

, containing the crankcase intake throttle valve

63

. Typical crankcase scavenged two-stroke engines use the pre-mix lubrication system, in which the oil is mixed with the fuel. As the air/fuel/oil mixture circulates into the crankcase, it provides the required amounts of lubrication to the sliding surfaces of the engine. As the engine object of this invention is basically air-scavenged, lubricants are not present into the scavenging gases. To circumvent this problem, very small amounts of air/fuel/oil mixture can be introduced into the crankcase with the scavenging gases to assure the lubrication of the lower components of the engine, but yet allowing very minimal effect on the total hydrocarbon emissions of such small amounts of raw gases escaping as scavenging losses. Engines where multi-position capabilities are not required, splash direct lubrication with oil recirculating system is typically used. It is obvious to the skilled in the art that the engine object of the present invention, offers unique opportunities for lubricating the internal components of the engine not offered by prior art two-cycle air-scavenged engines, such as piston timed air/fuel/oil mixture bleed into the crankcase.

Thus, from the foregoing description it should be readily apparent that the described embodiments of the invention provide a sound method of reducing unwanted emissions released into the atmosphere while still preserving the preferred characteristics of conventional two-cycle engines, which allows its use on portable equipment or applications where cost, package size, weight and emissions are the mandatory factors.

Of course, the foregoing description is that of preferred embodiments of the invention and various changes and modifications may be made without departing from the spirit and scope of the invention, as defined by the appended claims.