Engine valve means and porting

Abstract

A two-cycle crankcase compression internal combustion engine having extended and specially positioned intake porting and reed-type intake valves, with the porting and valves arranged to improve various of the operating characteristics of the engine.

Description

The present application is a division of my application Ser. No. 375,065, filed June 29, 1973, porting of the type herein disclosed was tested. As can be seen from the graph of test 3, power output below 5000 RPM is significantly increased, somewhere in the order of 50% and peak power of about 29 horsepower was achieved at a speed of about 7300 RPM. Moreover, power output beyond peak power speed falls off more gradually than that for the number 1 and number 2 tests. Further, maximum engine speeds of almost 11,000 RPM were achieved.

In test 4, an engine with the same modifications as that in number 3 and further including a carburetor with a larger venturi diameter, auxiliary transfer porting, modified exhaust porting, and a modified exhaust expansion chamber was used. Peak power on this engine rose to 34 horsepower at a speed of about 9200 RPM. Maximum engine speed was found to be in excess of 12,500 RPM.

It should be understood in connection with the vented reed valve arrangement heretofore discussed that valve assemblies having more than two reeds are contemplated according to the invention. For example, triple reed valve arrangements may be employed, in which event the secondary reeds will be apertured or vented, so that the third or tertiary reeds serve to open and close the vent in the secondary reeds. In this case, the tertiary reeds are desirably smaller and more flexible than the secondary reeds.

With further reference to FIG. 3, it is pointed out that the arrangement of FIG. 3 is similar to that described above and that all of the parts described above are also employed, but the reed valve assembly is differently positioned in the valve housing 20. In effect, the reed valve assembly in FIG. 3 is merely rotated 90° in the valve housing, as compared with its' position in FIGS. 1 and 2. Because of the angular position of the valve assembly in FIG. 3, the reed valves themselves occupy a different position in relation to the porting and to the axis of the cylinder. This is of advantage because it allows a more predictable flow pattern of combustion fluid through the system, expecially during high speed engine operation. The flow is more evenly divided between the sets of reeds disposed on each side of the valve body 29 and is directed in a manner which conforms to the natural directions of the fluid flow through the engine, i.e., curved toward the sides of the crankcase where the transfer passages are open to the crankcase, by reason of the fluid flowing through the valve port which acts as an orifice having a fixed side and a yieldable side defined by the reeds which cause the flow to curve. By reason of the orientation of the valve assembly as shown in FIG. 3, the reeds are placed closer to the port 46 and the flow into port 46 is smoother because the fluid does not have to flow upwardly from the reeds as in the FIG. 1 embodiment.

This orientation of the reed valves is also desirable because it improves cold starting of the engine, which is particularly advantageous with engines requiring manual starting. The reason for this is that when the engine is at rest, there is no vacuum in the crankcase to operate the reeds. Thus, at start-up, the engine must be cranked to develop sufficient vacuum in the crankcase to cause the reeds to open and allow the passage of the combustion fluid into the engine. When the valves are positioned as shown in FIG. 1, the combustion fluid passing through the bottom set of reeds must flow upwardly as it enters the valve assembly 27 so that it can exit through the vent in the primary reed 38. These factors require a greater vacuum to be developed in the crankcase in order to draw the combustion fluid through the valve assembly and this necessitates higher cranking speed at start-up. When the arrangement shown in FIG. 3 is used, the only forces which must be overcome in order to draw combustion fluid through the valve assembly are the resistive forces developed by the secondary reeds. This reduces the vacuum required to draw the combustion fluid through the valve and correspondingly decreases the cranking speed or starting effort required. In some engines this difference is so great as to make it practical to manually start the engine if the valve arrangement of the invention is used, whereas manual starting would not be practical without the valve arrangement of the invention.

In connection with the orientation with the reed valves as shown in FIG. 3, it should be kept in mind that in many installations such as motorcycles and snowmobiles, the intake passage and also the engine itself is somewhat inclined in a direction so that liquid fuel would tend to flow from the carburetor through the intake passage and intake port into the cylinder. This inclination is shown in FIG. 3. With the valves oriented as in FIG. 3, some liquid fuel may readily leak past the valves or may accumulate immediately upstream of the valves, which is in contrast with the condition when the orientation of the valves is as shown in FIGS. 1 and 2. The arrangement of FIG. 3, particularly where the intake passage and engine is inclined, is therefore of special advantage where easy starting is an important factor.

It should be noted that the foregoing benefits are achieved from orientation of the valve assemblies as shown in FIG. 3 with vented reeds as heretofore disclosed, and also with single reed designs. Single reed designs benefit because the single reeds are less yieldable than the vented reeds and therefore do not open as easily.

Turning now to FIGS. 4, 5, and 6A to 6E, attention is directed to the porting employed in accordance with the present invention.

FIG. 4 shows a cross-section, taken along line 4--4 of FIG. 3, of a typical cylinder showing the preferred manner of mounting the valve assemblies 27 in relation to the cylinder. In this embodiment, two valve assemblies 27, are positioned vertically as discussed above with respect to FIG. 3 in a housing 20 which is attached to the cylinder C. The valve assemblies 27 are positioned so that each valve assembly is aligned with one of the intake ports 18. The use of the two valve assemblies 27 is advantageous because it causes the flow of combustion fluid to agree with the natural flow pattern of the engine, and this results in a smoother, more directional flow of combustion fluid through the engine. The combustion fluid flows through each of the valves 27 into an aligned inlet port 18 and into the crankcase of the engine and from there into the transfer passages 53 and is introduced to the combustion side of the piston through the transfer ports 16. It should be noted that this is particularly important in the preferred embodiment of cylinder arrangements as shown in FIGS. 3 and 4 because in such designs, the axes of the transfer ports 16 are angularly displaced from the axis of the intake tract by about 90°, as is shown in FIG. 4. These designs require that the combustion fluid make an abrupt change in direction once inside the engine, i.e., a direction change of 90° to either side of the engine to enter the transfer passages 53, and without the arrangements as shown in FIGS. 3 and 4, the charge has little inherent directional tendency to flow toward either side of the crankcase, and in fact does not do so until after the charge has been compressed and flow starts through the transfer passages. At high operating speeds, the time during which the change in direction can occur is short and therefore the change must occur quite rapidly. When using the double valve assembly arrangement shown, the valve assemblies give direction to the combustion fluid streams before the streams enter the engine. This predetermining results in the combustion fluid stream undergoing the direction change more efficiently, rapidly, and smoothly, thereby ultimately resulting in the delivery of a larger volume charge to the combustion side of the piston.

It should be noted that the axis of the entire inlet tract of the engine shown in FIG. 4, including the valve assemblies 27 and the intake ports 18, is positioned so that it is aligned along a radius emanating from the center of the cylinder bore and is angularly offset from the axes of the transfer ports.

There is shown in FIG. 5 a piston 56 which is usable in conjunction with a cylinder such as shown in FIG. 4. The piston 56 includes two spaced piston ports 58, each of which is positioned to be aligned with one of the intake ports 18 of cylinder C. An important aspect of the piston shown in FIG. 5 is that the height of the ports 58 is greatly increased over that of prior designs. As will be hereinafter more fully explained, this results in allowing the inlet ports 18 (and thus the inlet tract including the reed valves) to communicate with the crankcase at all times during the engine cycle. The height of the ports 58 can be increased in designs as shown in FIGS. 3 and 4 both with single reed valves and with vented valves as heretofore disclosed. This is so because the improved fluid flow resulting from the vertical placement causes the engine to start more easily and does not require the intake ports to be closed off from the crankcase in order that sufficient vacuum be developed to operate the reeds--the mode of operation necessary in designs having horizontally oriented reeds. The increased height of the piston ports, and the concomitant increase in port area allows a longer induction period and thus greater charge of fuel to be inducted into the engine and this results in higher engine outputs.

There is shown in FIGS. 6A-E a schematic representation of the operating cycle of an engine employing vented reed valves of the type heretofore described and employing a ported piston as shown in FIG. 5.

FIG. 6A shows the position of the piston 56 just slightly before it reaches bottom dead center. The combustion fluid charge compressed by the descending piston 56 has exited from the crankcase 60 and is introduced via the transfer ports 16 and somewhat through auxiliary port 46 to the combustion side of the piston. As described above, the rapid exiting of the compressed combustion fluid from the crankcase 60 causes a vacuum to be created in its wake in the crankcase 60. This vacuum is transmitted via the piston port 58 to the reed valves which open allowing the introduction of additional charge of combustion fluid through the auxiliary transfer port 46 to the combustion side of the piston and also into the crankcase 60 through the piston port 58 resulting in extended delivery of charge through ports 16 (as shown by the arrows). This creates what is known as a supercharging effect in the lower RPM ranges and results in higher engine outputs.

There is also a supercharging effect which occurs at high RPM. At high RPM, it will be recalled, the secondary reeds remain open, by reason of the fact that the incoming charge of air-fuel mixture is travelling at high velocity and has significant momentum and, therefore, the charge continues to flow through the open vent in the primary reed and allows the system to maintain a higher delivery rate. Also, at piston bottom dead center, typical exhaust expansion chambers are supplying suction to the cylinder. At this position, the auxiliary transfer port 46 is open to the combustion side of the piston, and because of the high momentum of the incoming charge which maintains secondary reeds open at high RPM, a portion of air-fuel mixture flows directly from the intake tract, through the port 46. Thus, this portion of the charge bypasses the crankcase at high RPM, to fill the cylinder thoroughly. Any of the charge tending to escape through the exhaust port as the piston moves upwardly is now held in the cylinder by a positive reflective wave generated by a typical exhaust expansion chamber. Also, because the secondary reeds remain open at high RPM ranges, there is an increase in the amount of charge drawn into the crankcase through the skirt port 58 and a correspondingly increased flow of gases through the crankcase and to the transfer ports.

FIG. 6B shows the piston as it has just closed off the transfer port 16 and auxiliary port 46. The piston is ascending, thereby creating a vacuum in the crankcase which is communicated to the reed valves via the piston port 58, thereby causing the reed valves to open further. Combustion fluid flows from the inlet port 18, and also from the auxiliary transfer port 46 when the piston 56 has moved high enough, through the piston port 58 to the crankcase. At this point, the skirt of the piston has not yet begun opening the intake port 18.

In FIG. 6C, the bottom edge of the skirt of the piston 56 has cleared the intake port. Under these conditions, the reed petals are open, combustion fluid flows beneath the bottom of the piston and also through the piston port 58 into the crankcase. It should be noted that combustion fluid flowing through auxiliary transfer port 46 is directed upwardly through the piston port 58 toward the underside of the top of piston 56. This latter flow cools the top portion of the piston.

As shown in FIG. 6D, the piston 56 is approaching the top of its stroke, the bottom edge of the skirt has completely opened the intake port 18, thereby allowing a great volume of combustion fluid to be drawn into the crankcase through the open reed valves.

FIG. 6E shows the piston 56 descending and closing off the intake port 18. The piston is of course compressing the volume of combustion fluid drawn into the crankcase during the previous up stroke of the piston. At this time, the pressure of the fluid in the crankcase is greater than the pressure on the upstream side of the valve assembly and blowback of the pressurized charge is prevented by the closed reeds in the lower RPM ranges and by the restricted area of the vents and the momentum of the incoming charge entering through the vents at high RPM ranges. It should be noted that at all times throughout the cyce cycle the crankcase is in communication with the intake tract, either via the piston port 58, the inlet port 18, or a combination of both. This provides for the induction of larger quantities of combustion fluid into the engine and results in higher power and higher torque outputs.

FIG. 7 is a graph based upon the type of valve and port arrangements shown in FIGS. 1, 3, 4, 5 and 6A to 6E, including the vertically extended porting 58 provided in the piston skirt as shown in FIG. 5. In FIG. 7 the graph there shown plots two curves, curve 1 representing typical operating behavior of the reed petals at low intake velocities, i.e., engine speeds below the power peak, and curve 2 representing typical operating behavior of the reed petals at high intake velocities, i.e., engine speeds above the power peak. As has been seen, at high intake air velocities, the intake is open to at least some extent throughout the entire cycle of operation of the engine. As plotted in the graph of FIG. 7, 180° represents the bottom dead center position of the engine crank, and 360° represents the top dead center position of the engine crank.

It will be noted that the vertical scale of the graph of FIG. 7 represents the degree of reed valve opening, graduated at quarterly intervals from zero opening to full opening, and the lowermost quarter of this scale comprehends the extent of opening provided by the secondary reeds, it being assumed that at the one-quarter position on the graph the secondary reeds are fully open.

The graph of FIG. 7 also shows that even at low intake air velocity, the duration of reed or valve opening is extended throughout approximately 240° of the cycle of operation. This aids in maintaining relatively high output and performance at low engine RPM, under which condition both the primary and secondary reeds cycle, as has been described.

The duration of reed opening as described above in relation to FIG. 7 is greater than prior arrangements both at low as well as at high intake air velocity, and these conditions can only be achieved when the skirt porting 58 is high enough to be open whenever the piston skirt would block communication from the intake passage to the crankcase. In prior arrangements, where the skirt port is closed during a portion of the cycle, the commencement of opening of the valve is delayed to or beyond the 180° position, i.e., bottom dead center. Such prior arrangements adversely affect the torque at both high and low engine RPM.

It should be noted that the graphs shown in FIG. 7 are representative of engines employing standard transfer port timing, i.e., usually not in excess of 120° duration. It has been found that when using vented reed valves as herein disclosed, especially in conjunction with piston porting as heretofore described, that the height of the transfer ports 16, as shown in FIGS. 1, 3, and 6A-E, 9, and 10, can be raised to give greater transfer duration. Engines having transfer port durations of about 148° have been found to have power curves as depicted by line 4 of FIG. 2G. It will be noted that greatly increased power results at high RPM. In addition, the height of the exhaust port can be raised, resulting in increased scavenging time and concomitant higher engine outputs.

Turning now to the graph of FIG. 8, the graph indicated by the numeral 1 represents a prior known single reed valve engine and and is characterized by rapid drop-off of horsepower after the power peak is passed. The curve identified by the numeral 2 is similar to curve 3 of FIG. 2G, and illustrates one arrangement or embodiment of the present invention incorporating a vented reed valve assembly. This curve shows much less tendency for the horsepower to drop off after the peak is reached. In another embodiment conforming with the present invention of the kind shown in FIGS. 3 and 4, in which multiple pairs of reed valves are arranged and in which a pair of spaced intake ports 18 are provided, a horsepower curve is shown by numeral 3 in FIG. 8 is secured. Here it will be seen that the peak horsepower is still higher and further, that the horsepower at the higher RPM levels off, instead of dropping sharply, as in the case of curve 1.

Turning now to FIGS. 9 and 10, there is here shown still another feature as applied to arrangements similar to those illustrated in FIGS. 3 and 4. Similar parts are again identified by the same reference numerals. In these figures, however, additional ports, herein referred to as "injector" ports, are provided. Two injector ports are illustrated at 62, 62. Each of these ports interconnects one of the intake passages 18 with one of the transfer passages 53, as is shown in FIGS. 9 and 10. These injector ports are open at all times, and serve to increase intake of fuel at the higher RPM's , especially above 6000 or 7000 RPM.

It will be noted from FIG. 9 that the longitudinal axis of the injector ports 62 is arranged at substantially a 90° angle to the axis of the transfer passage 53. When the charge contained in the crankcase is pressurized by the descending piston, the charge is caused to flow upwardly through the transfer passages 53 to the transfer ports 16 at high velocity. In accordance with Bernoulli's Principle, the rapidly moving charge in the passage 53 moving past the opening of injector port 62 causes an eductor effect in the injector port 62 which causes a low pressure to exist in the port 62, which low pressure is communicated to the intake tract just downstream of the reed assemblies. In this manner, a quantity of charge is drawn from the intake tract downstream from the valve assembly, through the port 62 and into the transfer passage 53. This results in a higher density charge passing through the portion of the transfer port between the injector port 62 and the transfer port 16. It is believed that injector ports can be used with beneficial results in two-cycle engine designs having valving in the inlet tract, for example, rotary intake valves. As will be noted below, in connection with the discussion of FIG. 11, especially good results are achieved when injector ports are used in engines having reed valves, especially vented reed valves of the type disclosed herein.

It is also preferred, as is shown in FIGS. 9 and 10, to provide a partition or wall 64 between the two intake channels 18 and the two pairs of reed valves, thereby aiding in directing the intake flow through the channels 18 and into the crankcase through the porting provided in the piston skirt.

Comparative analysis of a given engine of somewhat higher horsepower than that employed as the basis for the graphs of FIGS. 2G and 8, both with and without the injector ports gives horsepower curves such as shown in FIG. 11. Here curve 1 is a curve of an engine conforming with the arrangements of FIGS. 9 and 10 except for the omission of the injector ports, and curve 2 represents the same engine altered merely by adding the injector ports. It will be seen that the peak horsepower has been raised, and further, that the drop-off of horsepower after the peak is further reduced, which is important at high RPM.

With the foregoing embodiments in mind, it is here desired to point out certain additional advantages and desirable operating characteristics secured when employing not only the multiple reed valves herein disclosed, but also when employing various of the porting features described.

The employment of reed valves also makes possible extensive increase in the total cross-sectional area of the intake porting, as is disclosed herein, and still further makes possible considerable increase in the total time in the cycle during which the valves are open, both at low speed and at high speed. The employment of reed valves further makes possible extending the porting 58 in the piston skirt to the point where the intake tract is open to the crankcase when the transfer ports are open. The use of reed valves also enables the vertical extension of the piston skirt porting to a point such that the intake tract is constantly open to the crankcase throughout the entire cycle of operation of the engine.

It should be noted that many manufacturers of two-cycle engines have been reluctant to adopt reed valves as a means of controlling the flow of the charge to the cylinder. This is believed to be because prior reed valve designs have added to the complexity of the engine design compared with piston port intake systems, and have exhibited unsatisfactory service life, yet have yielded only modest benefits in terms of somewhat higher power output at low RPM. Applicant's vented reed design, alone and in combination with the porting arrangements herein disclosed, has on the other hand achieved very significant increases in power output, torque output, and power band width. It is believed that these improvements make the adoption of reed valves by engine manufacturers much more likely.

Claims

1. A variable speed two-cycle crankcase compression internal combustion engine comprising a cylinder, a piston movable in the cylinder, a fuel intake system including an intake tract at one side of the cylinder and having reed valve means for controlling the flow through the intake tract, and said intake system including porting in the cylinder wall and in the piston located at the same side of the cylinder as the intake tract, the intake tract, the reed valve means and the intake system porting in the cylinder wall all being located to confront the bottom dead center position of the piston, the piston and the cylinder and piston porting being located and proportioned axially of the cylinder to provide flow channel means of substantial cross sectional flow area between the intake tract and the crankcase of the engine throughout the entire cycle of the engine including the bottom dead center position of the piston and thereby provide uninterrupted fuel intake during the stroke of the piston from bottom dead center to top dead center position, the piston porting including at least one port opening and every such piston port opening being positioned axially of the piston so that substantially the entire flow area of the piston port opening communicates with the cylinder porting in bottom dead center position of the piston, and a fuel transfer system for transferring fuel from the crankcase of the engine to the combustion side of the piston including a transfer passage communicating with the crankcase and having a transfer port through the cylinder wall and axially positioned to be closed by the piston in the region of top dead center position of the piston, said transfer port being offset from the cylinder and piston intake porting circumferentially of the cylinder so that a fuel flow path is provided in the transfer system from the crankcase to the cylinder at the combustion side of the piston independently of the intake tract.