EXHAUST SYSTEM FOR INTERNAL COMBUSTION ENGINES

Abstract: Exhaust system for internal combustion engines whose purpose is to take advantage of the energy normally wasted by the engine, in order to create vacuum inside the pipes (1) at the moment the exhaust valve opens. Vacuum which will suck out the gases from the cylincer, diminish engine temperature, reduce noises and improve engine efficiency. The elimination of all exhaust pressures is accomplished by means of a turbine (2), operating on the principle of pressure wave decomposition.

Description: Exhaust system for internal combustion engines. It is well known the low efficiency of internal combustion engines. It is known too that part of the energy is thrown away by the exhaust. The present invention is that of an exhaust system whose pur- pose is to take advantage of the energy wasted by the exhaust. First let's assume that in a 4 stroke engine the exhaust stroke combines waste of energy and negative work interfering in efficiency and performance. Actually what happens is that the time available for the engine to get rid of it's exhaust gases is extremely short. Of course the time during which the exhaust valve opens is enough for the gases to exit, but there must be no back-presure on the cylinder while it is on it's up-travel which is the exhaust stroke itself. That is why the exhaust stroke always begins at the final stages of the expansion stroke, and thus it is easy to conclude that the negative work is on one side while the loss is on the other. Furthermore the time of self exhaustion or the energy which expels the exhaust gases from the cylinder varies according to acceleration .or pressure on the cylinder while the time during which the exhaust valve opens varies according to engine rpm. The solution adopted in order to let an engine reach high rpm and consequently more power output is by an earlier opening of exhaust valves which cuts the expansion stroke, but relieves the piston from back—pressure on it's up-travel. The result of using such a technique is high fuel consump¬ tion and poor torque, balanced by high power output and rpm. The question remains unsolved. If exhaust valve opens early, say 50 ρ BBC some energy will be lost on the final expansion stroke but the engine will be able to reach high rpm since the time available for exhaust will be longer'.On the other hand if exhaust valve opens near EC the expansion stroke will be com¬ pleted but the engine will face heavy piston back-pressure at high rpm and thus poor horsepower output. Another factor which limits performance in I.C. engines is the exhaust hack-pressure, which controls the exhaust flow con¬ stricting it in order to smooth engine operation. Let's not forget the major enemy the atmospheric pressure. The object of the present invention is to provide a means of utilizing the energy wasted in order to per orm a new and useful role in internal combustion engines not using exhaust driven turbo-chargers. The purpose of the present invention is to take advantage of the energy wasted and use it to diminish to a minimum level the negative work and the losses occuring on exhaust stroke. Yet another object of the present invention is to provide a means of both controlling the flow and fighting atmospheric pressure, eliminating the need for exhaust back-pressure while damping noises. Among the several benefits of using such a system it can be said that the exhaust back—pressure will be eliminated due to the new constriction-free means of controlling the flow. The early opening of exhaust valves will be forgotten because back¬ pressure and atmospheric pressure will not hinder exhaust flow anymore. In addition the vacuum produced in the final stages of ex¬ haust stroke will be greater and continuous all through the rpm band. This will benefit strongly the scavenging effect during overlap, making possible smaller time for this low efficiency task. As the extra vacuum will be continuous and there will be no more back—pressure pile-up as engine acceleration and speed increases,it is possible for the compression ratio to be raised. The volume of residual exhaust gases left on the clearance will be smaller and the charge temperature consequently lower, de¬ creasing the ocurrence of pre-ignition and detonation while im¬ proving volumetric efficiency. The system comprises a turbine which is responsible for the energy transformation. It works connected to the exit end of the exhaust pipes (tubes) serving each cylinder of the engine. The channels of the turbine work as a movable extension of the pipes, receiving pulses and creating vacuum by pressure wave decomposition. The only exception is for the one cylinder engine where the turbine channels cannot share different pipes and thus it is a different operation. The most efficient exhaust Systems use single pipe per cyl¬ inder because it is important to direct the blast of the exhaust through a tube in order to create vacuuπ; which will suck out the remaining gases in the cylinder. The drawback of single pipe per cylinder is that the positive pulse can be con¬ verted into a negative one (oposite direction) which sometimes hits the positive and blocks the elimination of exhaust gases. Furthermore single pipe per cylinder must be tuned according to the rpm at which it will work. Outside that range the effect will be negative, making engine operation troublesome. The present invention proposes a method where the exhaust will be tuned to the maximum acceleration desired no matter what the rpm, and for other situations the efficiency will be diminished but never becoming negative. Tuning of single pipe per cylinder exhaust system is accom¬ plished by pipe lenght. The tuning of the new exhaust system differs from that of ordinary exhaust systems. First of all the cross section area transverse to the flow (here called C.S.A. ) must be kept unchanged all through the system from exhaust port to the end of the turbine channels. This is important since the energy will be taken mainly from the gases speed and any en- largement or obstruction represents loss. The curves on the pipes must be as open as possible and always near the cylinder , avoiding those near the turbine except if they are projected against the turbine blades in order to help it's driving. The pipes must be arranged so that they work on identical temperatures because it's internal temperature affects the tuning. The lenght of the pipes depends upon the maximum accel¬ eration desired with both engine and pipes hot. The volume inside each pipe will be aproximately 2 to 3 times that of the cylinder. As previously mentioned the pipe lenght tuning depends on it's internal temperature but other factors must be taken into account such as the turbine resistance to gas flow, i.e. hether it is driven by the crankshaft or by the exhaust jets or if the system comprises a catalytic converter or mufflers etc. The sum of C.3.Λ. for each exhaust pipe of the engine must be equal to the sum of C.S.A. for each turbine channel. As the number of channels must be at least double that of pipes con¬ nected to it, the C.S.A. of the channel in this case will be half that of the pipe. The exhaust valve as described above need not be opened early in order to achieve high engine rpm and power output. The overlap time will be narrower not only by compression ratio rise, but by a continuous and steady exhaust generated vacuum. The inlet valve closing can be retarded and this will benefit torque at high speed whithout affecting engine behavior at low speed (depending of course on the intake manifolds layout). The turbine rpm is crucial and must be kept to a minimum according to the operation cf the engine. For calculating the minimum rpm of the turbine or a given situation these things must be considered: the engine rpm, the pipe lenght, and the channel lenght which are the most important among others. For instance let's take a 6 cylinder engine with 6 pipe ends con- nected to a 12 channel turbine, the engine is running at 3,000 rpm at full throtle. The pipe is 1.6 meter in lenght while the channel is 20 cm. In this case we have 1 exhaust pulse every 12C 5 of crankshaft revolution. As the channel lenght is l/δ of the pipe lenght it would be necessary 8/6 of turbine revolution to cover the pipe lenght every 120 δ of crankshaft revolution. So at 3,000 rpm the turbine should be spinning at aproximately 12,000 rpm. The general formula for calculating the minimum rpm of the turbine is: TE = EE x pl/cl x ? KP where E? - the engine rpm TE - the turbine rpm pi .and cl — pipe lenght and channel lenght KP — number of pipe ends connected to the turbine ? — to be determined by tests This formula of course is or a rough calculation of the minimum turbine rpm and it doesn't take into account all the variable elements or a precise calculation such as accelera¬ tion. So it serves to ilustrate the relationship between dimen¬ sions for the exhaust system, while establishing a minimum basis of turbine rpm which only real tests on the various kinds of engine can demonstrate. The pipes must be arranged so that where it is connected with the turbine, the flow order or firing order describes a sequential circular movement in the same direction as that of the turbine. A lubricating system is necessary for the bearings of the turbine. In order to improve the energy exchange of the turbine (pressure x vacuum) the pipes can be divided so that the pipe end connections with the turbine will be twice as many and the frequency of pulses on the turbine will double, making possible the use of smaller turbine channel or lower turbine rpm. This solution somehow creates a difficulty in manufacturing since the number of turbine blades will double also. It can be useful for bi-cylindrical engines. The best option for the variable operation car engine is to connect the turbine to the crankshaft by a shaft so that the turbine will rotate in a ratio determined by the engine rpm. It is important to keep the turbine running during deceleration to avoid engine malfunctioning caused by uncontrolled gas flow at such conditions. As the speed of the exhaust will increase, the heat con¬ tained on the gases will be carried quickly from the valve port to the turbine where it will concentrate. As a result the tem¬ perature in the exhaust valves and cylinders will decrease but the latter will be partially compensated by the delayed opening of exhaust valves. The drawings on the accompanying sheets represents the in- vention and it's embodiments which are as follows: Fig. 1 is a lateral view of the system showing the pipes(l) which extends from the valve ports to the connection pipe x turbine(2), the turbine case(3), and the driving shaft(4). Fig. 2 is a lateral perspective view of the turbine,showing it's blades(5), the channels(6) and the εhaft(7). The reference numbers 8 and 9 shows the inlet and outlet of the channel which can be extended if necessary. Fig. 3 shows the pipe ends which are to be connected to the turbine. This piece can be twisted in order to direct the jets against turbine blades so as to help it's driving. The connec¬ tion with the turbine forms a ring(lθ), and must be as close as possible to the channels. Fig. 4 is a cutaway view of the system and shows the shaft (11) of the turbine, it's bearings(l2) and the driving shaft (13) which will be connected to the crankshaft. The joints(l4) of the driving shaft transmit power through angles and elimi¬ nates vibrations. The turbine blades(l5) must work very close to the case(l6) so as to avoid undue passages of air. To con¬ nect the system with it's further embodiments it will be neces¬ sary a manifold(l7). Fig. 5 s a detailed view of the driving shaft joint which consists of a spring(lδ). Fig. 6 is a schematic diagram of flow in ordinary exhaust systems, showing how the gases are exhausted from the beginning of the exhaust stroke (usually 50 s bbdc) until 0 s after bottom (dead) center. The vertical lines inside the pipes are the re¬ presentation of pressure at that point, while the horizontals under each pipe represents the speed. The abrupt variation in speed and pressure on the first pipe is the sound wave.' Fig. 7 is the same diagram of flow or the invention, where the gases will not face back-pressure and consequently will be carried much faster and efficiently. It can be seen the sound wave being damped on the first pipe by simply traveling through a very low pressure medium. The efficiency of the sound damping will be near to that of the exhaust. As the exhaust will be tuned for the maximum acceleration desired, the maximum vacuum inside the pipes will be achieved at such conditions, and so the sound damping will be best at the exact point where noises will be higher. It can be seen in Fig. 7 that when the high¬ speed gases hit the turbine they will be braked, and the resul¬ tant pressure wave must not reach the cylinder while the valve is open. Fig. 8 is a graphic of pressure x lenght which explains the principles of pressure wave decomposition. The pipe(20) and the channels(24) only demonstrate where is the connection of pipes with the channels and have no other purpose except this. The pressure wave(1 ) is that which hits the turbine and have to be decomposed in order to generate vacuum. The waves number (21) and (22) could be produced by the short channel but the higher and more efficient was made by a turbine with a higher ratio of channels to pipe end connections. To produce a high efficiency wave (23) with a low ratio of channels to pipe end connections, it would be necessary a long channel and consequently a heavier turbine. The vacuum waves(25) were produced by the waves(21,22) and will be thrown into the next pipes to receive exhaust gεn- erated pressure waves.

Claims: Claims : 1- Exhaust system for internal combustion engines,comprising a turbine which is connected to the exit end of the exhaust pipes, where the said pipes work for each cylinder of the en¬ gine, being in a number equal to them. 2- Exhaust system for internal combustion engines, as set forth in claim 1, comprising a turbine which is divided into channels which work as movable extensions of the pipes, being the number of these channels at least double that of the pipe ends connected to them, and being the number of pipe ends equal to that of the cylinders or double that number if the pipes divide before connecting to the turbine. 3- Exhaust system for internal combustion engines, as set forth in claim 1 and 2, comprising a rotating turbine working inside it's case which has the shape of a volcano, being this case composed of 2 openings: the smaller one or the inlet where the exit end of the pipes are connected to the channels, and the larger one or outlet where the exhaust gases are freed to the atmosphere or further embodiments of the system. 4- Exhaust system for internal combustion engines, as set forth in claim 1 through 3, comprising a turbine which is a rotating wheel whose outer surface or circumference is covered with the channels working inside and very close to it's case, and where the rotational movement is provided by a shaft which passes through it's center, being this shaft the rotationε.1 structure of the turbine which can be connected to the crank¬ shaft in order to determine it's correct rpm. 5- Exhaust system for internal combustion engines, as se forth in claim 1 through 4, comprising a turbine whose connec¬ tion with the pipes has the shape of a ring, being this ring radialy divided so as to fit the exit end of the pipes (if seen from the turbine), or the inlet of the channels (if seen from the pipes), and being the vacant area in the center of the said ring, the space for the turbine shaft to extend to it's struc¬ ture.

Inventors: NIEMEYER ARMSTRONG FERNANDO AU, NIEMEYER ARMSTRONG, FERNANDO, AUGUSTO DE
Applicants: NIEMEYER ARMSTRONG FERNANDO AU, NIEMEYER ARMSTRONG, FERNANDO, AUGUSTO DE
Application year: 1990
Main classification: F01N5
IPC codes: F01D5, F01N5, F02B27, F02B29, F02B37