DE4241403A1 - Gas-generator engine assembly - keeps gas outlet temp. from expansion unit to min. by controlling gas inlet temp. and pressure - Google Patents

Abstract

The gas inlet temp. at the expansion unit (14) is the max. permissable for the material of the latter. This temp. and also the gas pressure, are such as to keep the temp. of the gas (32) leaving the unit outlet to the min. when the engine is running. The latter (16) can be four-stroke piston engine with inlet and exhaust ports, the latter ports being connected to the expansion unit inlet. Gas outlet temp. can be kept low by a pressure-charger (18) with charge-air cooling, connected to the engine inlet. The engine exhaust can be divided into high and low-pressure portions, feed respectively to high- and low-pressure stages in the expansion unit. USE/ADVANTAGE - Combines power of gas-turbine with thermal efficiency of diesel engine.

Description

The present invention relates to a new and ver improved engine, in both of the gas turbine and also advantageous from piston-cylinder engines Enter features.

Common engines for motor vehicles and entspre relatively low output power are petrol and diesel engines, both versions of the Piston / cylinder engine type. Although she relies on a relative High level have developed such engines various disadvantages. They are in terms of their off physically very large. Besides, they are Petrol or diesel fuel using piston engines in  big pollution units and continue for - Multi-substance applications are not well suited. These disadvantages are the consequences of the short combustion time and the basically unfavorable combustion chamber and shape a piston engine.

With the development and perfection of for the Application in ceramic heat exchangers adapted Material will be an opportunity for improvement thermal efficiency of heat engines he reichbar. However, this technology can be used for piston engine can not be used profitably. That already large output energy loss of piston engines is in a ceramic piston engine larger; every reduction the necessary engine cooling due to a higher Be operating temperature leads only to a nominal increase the braking power, but with a large increase in the Output energy loss.

The gas turbine engine was long as attractive mögli cher replacement for the motor vehicle engine considered. In addition to other benefits, the gas turbine is com pact, has favorable combustion characteristics and Can by the application of constantly improved high temperature material technologies, such as B. Keramiktechno logies, which greatly improves performance the. While the gas turbine engine has great success in terms On aircraft systems, he was referring to not successfully use automotive technology tet.

There are several reasons for this unsuitability. When first, a motor vehicle engine requires the ability the output power over a wide load range  provide. The base gas turbine engine has, if also a reasonable thermal efficiency Full load, a very low thermal efficiency at partial load. The resulting total fuel economy when used in a motor vehicle is thus bad.

Because of this problem, almost all are experimental Regenerative type gas turbine engine engines, the trying to get the lost output energy in one Regenerator or recuperator recover. The Fuel economy of such a motor is compared to a basic gas turbine better and in one Chen cases comparable to that of a piston engine. The Compactness of the base gas turbine is, however, by the Regenerator severely impaired. In addition, the Regenerator expensive and tends in some versions too Maintenance problems.

A more attractive alternative is a charge air cooled turbocharged gas turbine engine, as described in the US Patent 47 00 542 was disclosed. This arrangement uses the Intercooling and charging to the output sener directly reduce losses and thus the emergency for a regenerator. The Ab cut used for the intercooler a working medium with lower temperature than the regenerator and works with a smaller amount of heat exchange. He is thus cheaper, smaller and less due for maintenance problems as the regenerator or Recuperator. Although the use of a charge air cooler makes the system less compact than the base guest bine, it is still more compact than the regenerative one Gas turbine.  

During the intercooled supercharged gas turbines motor considerable possibilities for the replacement of against has used gas turbine engines, let others close the gas turbine's own characteristics, that they are suitable for low output applications, such as B. for motor vehicles, less suitable. The Degree of performance of the compressor and Turbi neneinheiten of the gas turbine engine depends on the Rey Nolds number of units. That means, ever klei ner the compressor and the turbine is, the larger it is the speed at which they must be operated. The Combination of high speed and the accordingly Precise Fer required for these small units Tolerances make them under the restrictions the current technology difficult and expensive in the constructive design and construction. Other disadvantages the gas turbines, such as the low acceleration capacity and the small specific power, also speak against the general application in motor vehicle engines.

A solution to the problems of shrinking from a gas turbine is the acquisition of a "compound engine" -nord voltage. It is recognized that the piston engine is very suitable net is relatively small volume at high pressures ken and temperatures while using a turbine by their high mechanical efficiency and large flow cross section is best suited to with large volumes at relatively low pressures hen. A combination of both elements in a row Arrangement seems to solve many of the existing problems eliminate individual units.

On the part of the turbine will be the problem of development of a system that deals with small volume, high  Speed and high temperatures bypasses; a piston engine can meet the requirements of small volume and easily handle high temperatures.

From the side of the piston, the ineffective gas is exhausted exchange process of the piston engine by the very high "Lei "boosting" the boosting (or turboaufla denden) attenuated compressor. In addition, the ho he exhaust energy loss as a result of the in the Ver Throttle motor available limited expansion the potentially much larger expansion of the working medium be overcome in the turbine. The economy comes especially to bear, if the "compound engine" on based on a charge air cooled power cycle, so that the working medium the largest possible expansion pressure ver achieved at maximum thermal efficiency. Existing known "compound engine" methods are based but not on the charge-air cooled power cycle. Despite their compelling logic, none of them is big technically successful.

Currently existing "compound engines" are generally of two types: the compound engine and the gas generator ratormotor. An example of the compound engine is the Napier E 145 Nomad aircraft engine, which requires a complex transmission to overcome its poor power-to-speed characteristics. The inflator motor is considered a more adaptable engine. A complex transmission is not required and the power-speed characteristic is excellent. A typical gas generator engine consists of a compression unit, a piston unit for combustion and control of the compression unit and a Nutzleistungsturbinenein unit. The compression "unit" and the piston "unit" may be indivisible as in the case of the free piston engine. In such a system, the expansion pressure ratio of the working fluid in the utility turbine is equal to the product of the compression pressure ratio in the compression unit and the ratio of the outlet pressure to the inlet pressure of the piston unit P _EX / P _IN . Thus, the maximum output power of the motor is dependent on the high P _EX / P _IN values achieved by the piston unit. As P _EX increases, so does T _EX , and accordingly the gas temperature peak in the piston / cylinder unit, so that the thermal efficiency of the gas generator motor is also dependent on P _EX .

Existing gas generator engines are either in the free piston design or in the on the two-stroke Kol benmotoren based design. For both embodiments, P _EX / P _IN must be less than one, because such engines rely on the incoming charge to scavenge or drive the burned charge out of the cylinder. A P _EX / P _IN ratio above unity would prevent the discharge of burnt charge through the exit slot. This requirement is a detriment to the output power and the thermal efficiency. In particular, such a gas generator engine is less efficient than the compound engine and the turbocharged diesel engine, especially at Teilla costs, and has no significant advantage over the diesel engine.

Accordingly, the purpose of the present invention is a Method of operating inflator engines create, which maximizes the thermal efficiency and accordingly gas generator engines in better burn  to transform the engines of excellence into the performance of the engine Gas turbine engine and the thermal efficiency of Combine diesel piston engines. Such engines can Meet the requirements of future drive tasks.

Another purpose of the present invention is Method of operating the piston / cylinder unit of the To create gas generator motor, where the outlet pressure of the piston / cylinder unit can be raised so that the corresponding outlet temperature on a thermal efficiency maximizing level kept can be.

Yet another purpose of the present invention is to provide procedures that are the basic principle of Charge air cooled-charged power cycle on the Case of inflator motors applies.

The present invention, referred to as a charge air cooled gas generator / expansion engine ("IGGEX"), ent holds one with a power turbine or one other expansion machine connected intercooler Inflator. The charge-air-cooled gas generator shapes Fuel and air in a high temperature and high and can change the shape of a piston / cylinder Accept unity. The power turbine forms the High temperature and high pressure charge in usable brake performance and an exiting low-temperature charge with atmospheric pressure. The two units are pneumatically interconnected. The temperature of the Charge outlet from the expansion machine is through the degree of intercooling and charging gesteu The maximum temperature of the cargo within the col ben / cylinder unit is characterized by the piston / cylinder off  Laßdruck controlled by the operational Adjustment of the intercooled gas generator and the Power turbine is kept at a high value. To show how the maximum temperature of the charge is through Lifting the piston / cylinder outlet pressure who received can, embodiments are presented.

In a first embodiment of the invention drives Four-stroke piston engine a charge air cooled compressor on, with the outflowing charge of the piston engine by a power turbine, preferably in two Stages, expands to produce the labor tax.

In a second embodiment, the four-stroke col gas cylinder replaced by two two-stroke pistons / cylinders, one of which is the intake stroke and the compression stroke while the other is working and off Stoßhub the four-stroke operation performs. A transmission valve between the piston / cylinder pair transmits the Charge from the first cylinder to the second cylinder. This allows the requirements of the suction and compression stroke of those of the working and off shock stroke be separated so that each piston corre sponding can be optimized.

As a variant to a 1: 1 ratio between An suction / compression and working / ejection cylinders a variety of suction / compression cylinders through an output trunk to be interconnected to deliver the cargo to a separate chamber, in the one continuous or continuously pulsating Burning is maintained before the Working / ejection cylinders is supplied.  

Both the degree of lowering of the temperature of the charge when exiting the expander and the Level of raising the maximum temperature of the cargo in inside the piston / cylinder use a common Characteristic in its effects - simultaneous improvement conditions of thermal efficiency and Ausgangslei ratio / weight ratio of IGGEX. Since the thermi efficiency especially for fuel economy but also increases the workload, and the output power / weight ratio above all else Contributes to the work, but also the fuel increased economy in drive applications have their simultaneous improvements are magnifying Effect. The present invention thus becomes wait for her to be an extraordinary fuel host efficiency and work performance.

A fuller understanding of the present invention and Their essential features may be taken into account the following detailed description of preferred Illustrative embodiments in conjunction with the attached drawings are obtained.

Fig. 1 is a block diagram of the components of the prior invention;

Figs. 2A and 2B are plots of the thermal efficiency of IGGEX as a function of supercharging pressure at various charge temperatures;

Fig. 3 is an illustration of the connecting part between the piston and the expansion elements, showing a double Kolbenauslaßventilanordnung;

Fig. 4 is an alternative exhaust valve arrangement for the embodiment in Fig. 3;

Fig. 5 is an illustration of a modified variant of the embodiment in Fig. 4, in which a nearly stoichiometric fuel mixture is used in the piston unit;

Fig. 5A is a temperature-entropy diagram for the embodiment of Fig. 5;

Fig. 6 is an illustration of an embodiment of the invention with two cylinders, in which the functions of a single four-stroke cylinder are shared between two cylinders;

Fig. 7 is a timing chart for the cylinder assembly of Fig. 6;

Fig. 8 is an embodiment of the present invention considering continuous combustion.

Fig. 1 illustrates the generalized arrangement of the components of an IGGEX engine 10. A charge air cooled gas generator unit 12 provides a high temperature and pressure input to an expansion unit 14 to which it is connected. The gas-generator 12 may standardize preferably a piston engine unit included 16 which is mechanically connected by a shaft 20 with a compressor or a charger 18th The compressor 18 may include commonly known intercoolers. The output from the piston unit 16 driven by the compressor unit 18 is a flow of a working fluid 22 with an increased pressure P _IN , which is merged with the fuel 24 and in the cylinders or combustion chambers of the Kolbenmo gate unit 16 is ignited to an outlet charge 26 with an elevated temperature T _EX and a pressure P _EX . This charge is performed by the device 14 Expansionseinrich, which usually consists of a useful power turbine 28 , which generates an output power to the shaft 30 . The exhaust 32 of the turbine 28 is returned to the atmosphere at atmospheric pressure.

As stated above, the piston outlet pressure in be piston engines were chosen to be slightly is less than the piston inlet pressure. This little one Difference serves to get the burnt gas out of the piston to flush the engine cylinder. When using the piston motors as a gas generator will be the matching utility power turbine selected so that their pressure ratio as a function of the flow rate characteristic Inlet pressure assumes that equal to this piston outlet pressure is.

The piston engine according to the invention is preferably designed as a four-stroke engine, whose properties have been selected to obtain an optimal P _EX value for driving the power turbine 28 . A Nutzlei stungsturbine can be selected so that their pressure ratio (inlet pressure as a function of the outlet pressure) as a function of the flow rate characteristic is such that the inlet pressure of the turbine, the laßdruck with the Kolbenaus equal to the boost pressure in the piston cylinders before the Auslaßventilöffnung (EVO ) and thus the relaxation loss of the piston / cylinder unit is eliminated. However, such a choice is not the most advantageous outlet pressure for the best thermal efficiency. In any case, the choice of Kolbenauslaßdruckes can not be performed independently of the selection of the Kolbeneingangs- (or charging) pressure. This matter is not obvious and will be further investigated below.

First, consider abstract designs of the IGGEX at different supercharging ratios R _SUP (ie, the ratio of supercharging pressure to inlet pressure (usually atmospheric pressure)) with temperature T _EX maintained at a constant level. Starting with a very low R _{SUP design} , to maintain the T _EX temperature requires a significant fuel-to-air ratio during operation, even when there is a low demand for recharge performance. This would lead to an excess of piston power beyond the power necessary to drive the supercharger unless the piston back pressure (the pressure in the off / intake charge line) has been greatly increased - higher than the boost pressure immediately before the exhaust valve opening (EVO ). At very low R _SUP is not just no relaxation or pressure drop between the boost pressure immediately before EVO and P _EX (turbine inlet pressure), but actually a return flow through the Kolbenauslaßventil. The return flow is of course a form of valve flow loss.

With the gradual increase in R _SUP , the return flow will decrease and at a specific R _SUP , depending on the assumed charge temperature T _EX , the inflator motor will have neither backflow nor expansion via the exhaust valve.

For designs with larger R _SUP , the power balance between the piston unit and the precompressors will require the pressure P _EX at a level which causes the expansion via the exhaust valve. As the R _SUP grows larger and larger, more relaxation is needed to lower the piston back pressure and meet the increasing and decreasing compression performance requirements. Referring now to Figure 2A, there is shown a schematic diagram of the thermal efficiency as a function of R _SUP for various selected T _EX values, in which the regions of backflow at low R _SUP and the regions of relaxation at high R _{SUP are} indicated.

Each continuous characteristic embodies a representation of the thermal efficiency as a function of the piston P _IN for a given T _EX with rising T _EX values plotted in the figure from bottom left to top right (T _EX ) ₁ <(T _EX ) ₂ <(T _EX ) ₃ are shown. Each curve has a maximum value for the thermal efficiency, represented by the broken circle, according to the basic principle of the charge-air cooled, supercharged power cycle for a corresponding Ph optimum. The location of the points of zero flow loss (neither return flow nor relaxation) on the curves is represented by the dotted curve. This dotted curve shows a monotonically increasing thermal efficiency compared to P _IN . This is explained by the fact that T _{EX is} not constant but increasing on the dotted curve. Also shown is the characteristic shift in the points of zero valve flow loss to higher relative P _IN levels as T _EX increases. By "higher relative P _IN " it is meant that the zero loss points are lower than the point of the P _IN optimum at low T _EX values and increasing towards the P _IN optimum and finally to the right side at high T _EX values of the P _IN optimum. This characteristic shift can be expected if the rise in P _IN at a zero valve flow loss as a function of T _{EX is} faster than the increase in the P _IN peak at the maximum thermal efficiency as a function of T _EX .

For the case of a "low" T _EX value, the possible design points for a given P _{IN are} considered equal to the P _IN optimum according to the selected T _EX value. With the rise of Kolbenauslaßdruc kes, and thus the Kolbenauslaßtemperatur, not only increases the output power, but also the thermal efficiency due to the reduction of the relaxation loss. The design point is accordingly set when the piston output pressure reaches its optimum value at the point where the piston outlet temperature is equal to T _EX - the temperature limit of the material. The thermal efficiency at the design point is an absolute maximum at the Ph optimum achieved by the charge air charge charging and at the optimum piston discharge pressure achieved by the means for increasing the piston discharge pressure. In this case, the design point includes a relaxation loss and a gas leakage at the expansion device. Still, the composite loss is less than any other possible combination of charge air cooler charge pressure and piston discharge pressure around the delivery point around depending on the temperature limit of the material.

In the case of a "high" T _EX , the possible design points along the constant P _{IN are} again considered equal to the P _IN optimum, as shown in FIG. 2B. The increase in thermal efficiency is achieved as a result of reducing the relaxation loss by increasing the piston outlet pressure. It is interesting that the increase continues even if the relaxation loss is eliminated and backflow loss occurs due to the further increase in the piston outlet pressure. The design point is again reached when the piston outlet pressure reaches the P _EX optimum and the piston outlet temperature reaches the material limit, T _EX . The design point for the P _IN -Optimum and P _EX -Optimum again coincides with the maximum thermal efficiency, which is characterized by the smallest composite loss from the return flow and the gas exit loss of the expansion device (in this case due to too low a temperature) ,

To keep repeating, the P _IN optimum and the P _EX optimum are determined simultaneously. Until the P _IN optimum is known, one can use the above procedure to find the P _EX value that matches the design value of T _EX for a given P _IN . A repetition of the process for other P _IN values follows until the maximum operating point of the thermal efficiency is found. At this (design) point, P _{IN is} the P _IN optimum, and P _{EX is} the P _EX optimum.

Based on the above considerations, P _EX and therefore the charge pressure during the exhaust stroke is to be maintained at a higher value than the P _IN value. Since the intake process in the cylinder can not take place until the boost pressure within the cylinder drops to the inlet pressure, no overlap of the operation of the exhaust and intake valve or outflow through the intake valve can occur. Instead, the exhaust valve must close almost at top dead center. However, to reduce the piston pressure without adversely affecting the P _EX value for the turbine, an additional purge outlet valve is provided, as shown in FIG. 3. This valve opens simultaneously with the closing of the off laßventils 40 and is accessible to a charging line 44 , which is supplied at a pressure which is slightly lower than P _IN , and remains open until a sufficient de purging is effected. Thus, the ejection process includes the following steps:

• 1. The exhaust valve opens some crankcase degrees - before bottom dead center.
• 2. The exhaust valve closes immediately after the Piston reaches the top dead center.
• 3. Simultaneously with the closing of the exhaust valve opens the second Spülauslaßventil.
• 4. The charge purge continues when the A laßventil opens at a crank angle behind the top dead center. If the charge line has a pressure below P _IN , the exhaust gas escapes further out of the cylinder.
• 5. The purge outlet valve closes and the suction process starts.

The above process, which forms a pair of outlets at different pressures, accordingly requires that the power turbine or expander be formed as a two-stage turbine. The P _EX charge expands as it passes through the first stage in an intermediate charge line from P _EX to a pressure slightly below P _IN . The extended charge is merged as it flows into this with the additional charge mass directly from the Zy cylinder through the Spülauslaßventil. This composite charge is then passed through the second stage of the turbine, where it continues to relax from the intermediate pressure to the atmospheric pressure.

As shown in Fig. 3, the provided with a piston 36 cylinder 34 takes the pressurized einströ coming charge 22 with a pressure P _IN through the Einlaßven valve 38 in a conventional manner. The high-pressure and high-temperature exiting charge 26 is at P _EX , T _EX through the first Auslaßven valve 40 , which opens during the exhaust stroke and remains open until the piston 36 reaches the top dead center, ejected. The exiting charge 26 expands through the first turbine section 42 and exits into an intermediate charge line 44 between the two stages 42 and 46 of the turbine 28 . The intermediate charge line 44 is designed to operate at a pressure slightly below P _IN , which of course is below P _EX , and accordingly has a working pressure below that of P _EX . At top dead center closes the exhaust valve and the second purge exhaust valve 48 opens. The Spülauslaßventil 48 carries the remaining exiting charge in the Zwischenladelei device 44 , where it is merged with the exiting charge of the first portion 42 of the turbine 28 . The intermediate charge conduit 44 serves as an inlet to the second turbine stage 46 in which additional mechanical work is obtained from the exiting charge leaving the second stage 37 at atmospheric pressure.

In Fig. 3, the purge outlet valve 48 is shown as a "drain" valve which is similar in operation to the conventional intake and exhaust valves 38 , 40 of the cylinder. Such an embodiment requires relatively complex connections to the cylinder head and may require significant changes to the conventional cylinder head design to physically permit the placement of an additional valve.

Consequently, in Fig. 4, an alternative to the type is presented as a drain valve. As shown therein, a rotary valve or other type of valve 50 , which may be staggered from the cylinder 34 , operates together with a conventional exhaust valve 40 and is timed to discharge the discharged charge into the first stage 42 of the turbine 28 or in the lower pressure intermediate charging line 44 to lead. By a remote from the piston and the cylinder arrangement of the valve 50 , the considerable transformation of the cylinder head is avoided. Instead, only the total opening time for the exhaust valve 40 needs to be modified to accommodate the time taken to transport both parts of the exhausting charge.

If the supercharging pressure ratio (P _IN / P _atm ) is very high, the degree of compression of the piston engine must be kept low, imposed mainly by the maximum pressure limit of the cylinder block. A low degree of compression, combined with the use of charge intercooling in the supercharger, results in a low charge temperature at the end of the compression stroke. This excludes the application of the compression ignition combustion. Depending on the maximum permissible outlet temperature corresponding to the projected inlet temperature for the turbine, the maximum fuel / air ratio can be limited to values below the stoichiometric. Consequently, running with lean mixtures driving internal combustion engines with spark ignition, such. As a stratified combustion engine or a multi-channel injection engine, which are capable of more economical combustion than the well-known, preferably selected for inclusion as a piston engine unit of IGGEX.

However, spark-ignition engines that run with nearly stoichiometric mixtures can be accommodated for use in the IGGEX piston unit with the following modification. Such piston engines must be connected to "oversized" superchargers. This necessary change is described in FIGS . 5 and 5A, in which a multi-stage compressor with intercooling cooling is used as a supercharger.

The trained as a multi-stage unit Verdichterein unit 18 has compressor 104 , 106 , followed by heat exchanger 108 and radiator 110th

"Excess" air from the last (second) supercharger 106 in state A bypasses the last stage charge air cooler 110 and the piston unit 16 and instead enters a heat exchanger unit 112 surrounding the piston exhaust tube. As a result of the heat exchange, the bypassed air is heated from A to B and the flask exit from C to D and then cooled from C to E during flashing. The piston outlet is still led to the first inlet of the first expansion device 42 of the turbine 28 and after the partial cooling and Wei teren relaxation to the lower pressure on facing intermediate charge line 44 . The first ex pansion is represented by the route DF. It warmed the bypassed air in state B and the charge leaving state F in the first stage of the turbine binds to the piston scavenging exhaust gas in state E in the intermediate charge line 44 to form a mixture in state G. Additional mechanical work is gained from the expansion of the mixture from G to H in the second stage 46 of the turbine. The temperature limit of the Turbi nenmaterials sets the limit for the gas temperature in the state D. In this embodiment, the Kolbenaustrittstemperatur at C may be higher than the temperature limit of the turbine material, since the piston outlet is first used to heat the bypassed air before at D in Contact with the turbine is coming.

An alternative method of increasing P _EX to a higher level than P _IN with minimum return flow across the intake valve is to divide the four-stroke piston / cylinder into two two-stroke pistons / cylinders, one for intake and compression and the other for power and compression emissions. The intake / compression piston / cylinder can be provided with a very high "degree of compression" so as to keep the return flow through the intake valve to an insignificant level. In addition, the separate piston / cylinder design has other potential advantages.

As stated above, the inventive Four-stroke piston engine not both flow losses over the exhaust valve as well as losses due to an off Keep the temperature above or below the optimum. Such losses can only be minimized. In general mine the expansion stroke for the cylinders is too short, when the piston outlet temperature is lower than a fixed set exclusive value. Consequently, a lower back pressure with accompanying relaxation necessary to provide sufficient performance at one to produce optimal inlet pressure. The temperature of the corresponding power turbine outlet will be something be higher than the optimum. When the piston outlet cam is higher than the optimal value, the expan The cycle is too long and the turbine outlet temperature be too low.

In anticipation of the availability of ever higher outlet temperatures due to advances in material technology, the correction of excessive expansion stroke and stroke volume, which are inevitable in a four-stroke piston engine, may be readily met alternatively in a separate piston / cylinder design. because each cylinder can be optimized to perform the two strokes that it has to fulfill. As shown in FIG. 6, the cylinders 52 and 54 are connected by a transmission channel 56 . Your ent speaking pistons 58 and 60 are connected to a common crankshaft 62 . The first piston 58 in the so-called cold piston cylinder 52 performs the intake and compression strokes, receiving the intake air via the intake valve 64 , while the second piston 60 in the hot piston 54 performs the working and exhaust strokes, the charge at P _EX , T _EX is exhausted through the exhaust valve 66 . The hot piston is displaced along an angle R _a .

The transfer valve 68 , which operates between the pairs of pistons and controls the flow between the cylinders through the channel 56 , opens approximately at the top dead center of the piston travel of the hot piston 60 , which corresponds to the angle R _a before top dead center for the cold Kol ben 58 , The transfer valve 68 closes approximately at the top dead center of the cold piston 58 , in accordance with the R _a- advance beyond the top dead center of the hot piston 60th The compressed charge is expelled from the cold cylinder 52 and allowed to enter the hot cylinder 54 during this period. The combustion is introduced in the hot cylinder at a selected crank angle of rotation before closing the transfer valve by a spark plug 70 after supplying fuel through the injector 72 . The volume of the transmission channel 56 is kept as small as possible in order to obtain a sufficiently high degree of "compression" in the composite unit of the hot cylinder and the transmission channel. The combustion is thus carried out on the basis of a "shock" under a nearly constant quantity condition.

With the separation of the piston cylinder into two separate Piston, it is possible to use a turbocharger gas generator as also apply a mechanically charged gas generator the.

Fig. 7 shows the valve timing for a turboge loaded separate cylinder design. As shown therein, IVO indicates the intake valve opening; IVC - the inlet valve closure; EVO - the exhaust valve opening; EVC - the exhaust valve closure; TVO - the transfer valve opening; TVC - the transfer valve closure; IGN - the start of ignition. The characteristic values for the cylinder parameters and timing with respect to the crank angle of rotation of the hot piston are as follows:

Drilling the cold cylinder | 0.16 m Bore of the hot cylinder 0.12 m Stroke of the cold cylinder 0.12 m Stroke of the hot cylinder 0.12 m Length of the connecting rod of the cylinder 0.22 m "Compression degree" (ratio of volume at BDC to volume at TDC) of the cold cylinder 25 "Compression degree" of the hot cylinder / transmission channel 13 Cranking angle displacement of the hot relative to the cold cylinder R _a 50 °

R of IVO = 52 ° (2 degrees ATDC of the cold cylinder)
R of IVC = 245 ° (15 degrees ABDC of the cold cylinder)
R of TVO = 0 ° (50 degrees BTDC of the cold cylinder and TDC of the warm cylinder)
R of TVC = 53 ° (3 degrees ATDC of the cold cylinder and 53 degrees ATDC of the warm cylinder)
R of EVO = 140 ° (40 degrees BBDC of the hot cylinder)
R of EVC = 4 ° (4 degrees ATDC of the hot cylinder)
R of INJ = 47 °
R of IGN = 49 °

For a given design where the outlet temperature matches the material temperature limit, there is an optimum volume or bore ratio between the cylinders. Such a determination may be made on an experimental basis by the development of a series of isotherms P _IN as a function of P _EX / P _IN for a specific volume ratio. On the curve, in accordance with the desired T _EX, the optimum P _IN lies. In addition, isotherms for a given outlet temperature for embodiments with different volume ratios are then developed. If the series of isotherms is represented together, the highest curve agrees with the optimal volume ratio.

More specifically, in the above manner optimal volume ratio is correct with the maxima len power output match; it can not exactly match match the maximum thermal efficiency, because the engine block cooling of the various carburetors something can be different. In practice it can However, it can be accepted that the optimal volume ver the maximum performance and that optimal volume ratio with the maximum thermal Efficiency sufficiently matches, since no real Difference lead to the selection of one or the other becomes.  

The previously illustrated embodiment of both the individual as well as the separate cylinder use the combustion in a "jerky" process, less in a continuum process.

The intermittent nature of the combustion process of conventional piston engines is both the main reason for their economic success as well as the reason for their inherent limitations in reduction the pollutants generated by them.

Instead, as described above, the charge directly from To convert cold into the hot cylinder, one can Modification come to be used that a variety of contains cold cylinders passing through a manifold are interconnected to the charge in a sepa rate combustion chamber in which a continuous continuous or continuously pulsating combustion is maintained. This type of combustion process This was taken over in the original Brayton cycle men. Such a combustion process that has been over 100 years Years ago, quickly lost popularity Most of the Otto four-stroke internal combustion engine because of Possibility of achieving higher compression ratios and the reduction of pumping and friction losses in the the latter arrangement. With the application of a high Charge pressure in the present invention is a high (actual) degree of compression in the Ansaugver sealing cylinder no longer necessary. Besides, it tends the high inlet pressure, the pump and Reibverver lessening losses. In conjunction with IGGEX and his reduced disadvantages, the continuous Ver Achieve a cleaner combustion than one intermittent combustion in an intermittent process.  

Furthermore, a smoother running of the engine with advantage good fuel-efficiency and reduced War possible.

Such continuous combustion in the form of a "Wankel" or rotary piston engine using rotary lobes or rotors is of particular interest. This is shown in FIG. 8. As shown therein, a suction compressor rotor unit 74 may include a first pair of rotors 76 , 78 within a shell 80, termed a cold rotor housing. As is well known, the shape of the rotor allows a plurality of chambers or spaces, usually 3 in number, the volume change ver with the rotation of the rotor. The angular positions of the two rotors are staggered so that the compressed charges flowing therefrom are spatially separated to provide a more consistent and uninterruptedly connected discharge charge. A charged A gangsladung is introduced through the inlet port 82 in the Ro torkammern, the compacted exiting charges are guided through outlet check valves 86 to a manifold 84 and collected in it. The manifold 84 supplies the combustion chamber 88 into which fuel is also injected. The Burn drying products are transferred via valves 98 to the expansion ejection rotor unit 90, the rotors 92, has 94 in the casing 96 which are in turn arranged successively offset angularly with respect to create a continuous usable combustion chamber for the combustion emissions. Due to the expansion in the chambers formed by the rotors in the order sheathing, the rotors, the ver by the shaft 100 with the suction rotors 76 , 78 and the supercharger are connected, driven to operate the compressor and the compression rotors, the P _EX / T _EX - let out through outlet slots 102 to the turbine expansion unit.

Claims (13)

An improved inflator motor comprising an expansion device ( 14 ) having a high pressure inlet opening and a low pressure outlet opening having means connected to the inlet opening for producing a gas supply having such pressure and temperature characteristics (P _EX ; T _EX ) that the temperature is maximum by the material of the expansion device ( 14 ) is allowed and the pressure is selected so that the temperature of the outlet opening of the expander ( 14 ) leaving the exhaust gases ( 32 ) during the operation of the engine is mi mi nimated mi. 2. Apparatus according to claim 1, characterized in that the means for generating the gas supply from a four-stroke piston engine unit ( 16 ), the inlet and outlet ports and whose laßöffnung from the inlet opening of the Expansionsein direction ( 14 ) is connected. 3. Apparatus according to claim 2, characterized in that the means for generating the gas supply additionally consists of a means for minimizing the exhaust gas temperature of the outlet of the expansion device, which is connected to the inlet opening of the piston motor unit ( 16 ). 4. Apparatus according to claim 3, characterized in that the means for minimizing the exhaust gas temperature of the outlet of the expansion device ( 14 ) is a charge air cooled supercharger ( 18 ). 5. The device according to claim 2, characterized in that the gas generating agent gate unit of a means for dividing the exhaust gas of the Kolbenmo (16) is additionally in high and low pressure components and the expander (14) follow-reaching high and low pressure stages (42; 46 ), wherein the high-pressure stage ( 42 ) is connected to the inlet opening of the expansion device ( 14 ) and the low-pressure stage and the outlet of the high-pressure stage ( 42 ) connected to a second inlet port and the high and Niedrigdruckabgasantei le to the inlet opening of the expansion device ( 14 ) and the second inlet opening are guided. 6. The device according to claim 5, characterized in that the means for dividing the exhaust gas in high and low pressure parts of a piston cylinder ( 34 ) which has a first Auslaßventilmittel ( 40 ) which is adapted to during a before the lower Dead time beginning and ending at the top dead center time interval to be effective, wherein during this period, the high-pressure exhaust gas share is defined, and has a second exhaust valve medium, which is adapted to effective during a second, at top dead center and ending shortly thereafter, the time interval be, wherein during this period the low pressure exhaust gas portion is defined, and the first Auslaßven tilmittel ( 40 ) with the inlet opening of the ex pansionseinrichtung ( 14 ) and the second Ventilmit tel with the second inlet opening of the expansion device ( 14 ) is connected. 7. Apparatus according to claim 6, characterized in that the first and second Auslaßventilmittel ( 40 or 48 ) in the wall of the cylinder ( 34 ) is arranged. 8. The device according to claim 6, characterized in that the first and second Auslaßventilmittel ( 40 or 48 ) consists of a valve ( 50 ) which is arranged in an exhaust manifold for the piston motor unit ( 16 ) and adapted to alternatively to the Inlet port of the expansion device ( 14 ) and the second inlet port to connect, wherein the exhaust manifold is connected to the cylinder ( 34 ) by a Hauptauslaßventil ( 40 ). 9. Apparatus according to claim 2, characterized in that the four-stroke piston engine unit ( 16 ) is a externally ignited, homogeneous loading motor. 10. The device according to claim 2, characterized in that the means for minimizing the exhaust gas tempera ture of the outlet of the expansion device ( 14 ) is a charge air cooled supercharger ( 18 ) having a first outlet for delivering a compressed gas discharge at a first temperature and a second outlet for delivering a compressed gas discharge having a second, higher temperature be seated, wherein the second outlet is connected in a heat exchange relationship with the exhaust pipes, whereby the compressed, the second temperature having gas output is heated and the gas is cooled in the exhaust pipe, and the second outlet at the second inlet port ends. 11. The device according to claim 2, characterized in that the four-stroke piston motor unit ( 16 ) consists of series-connected cold and hot Kolbenein units ( 58 , 60 ). 12. The device according to claim 11, characterized in that the motor unit ( 16 ) additionally consists of a combustion chamber which is separate from the cold and hot pistons ( 58 ; 60 ), the power output of the hot piston ( 60 ) to drive the cold piston ( 58 ) is connected and a supercharger ( 18 ) for generating a charge Ganges for the motor unit ( 16 ). 13. The apparatus according to claim 2, characterized in that the four-stroke piston engine unit ( 16 ) is a rotary piston engine.