DE2035605A1 - Internal combustion engine - Google Patents

Description

PATENT LAWYERS

D R. I. M AAS

D R. W. PFEIrFER TR-17

DR-RVOiTHEMLEiTNER

8 M ü \ i CH HM 2 3
GEKiRSTR. 25 - Th'L 39 02 36

Treadwell Corporation New York, New York, V.St.A.

Internal combustion engine

The invention relates to a four-stroke piston internal combustion engine with a fuel inlet device.

The theoretical mechanical work generated by a heat engine corresponds to the difference between the the heat supplied to the machine and the heat dissipated from the machine. If a Carnot cycle or

- 2 009887 / U4S

any other reversible work cycle is applied, the greatest possible conversion of the available heat into mechanical work, i.e. the maximum thermal efficiency, a puncture of the heat supply _{temperature (T 1} ) and the heat dissipation temperature (T ;-), namely:

Thermal efficiency (fraction) = 1 m—. T ₁ and T ₉ are

I ₁ ι d

absolute temperatures. It is therefore apparent that a maximum efficiency is achieved when the heat supply temperature T as high as possible, and the heat dissipation temperature T is as low as possible. ₂ In an isentropic expansion process, the engine's inherent expansion ratio determines the ratio of the discharge temperature T ₂ to the supply temperature T ₁ . For a given value of T. the maximum thermal efficiency is therefore determined by this expansion ratio.

Attempts to improve the thermal efficiency of reciprocating internal combustion engines with a diesel or Otto duty cycle have mostly been directed to increasing the compression ratio, because this also increases the expansion ratio. As stated above, for a given, maximum allowable heat supply temperature, a larger expansion ratio leads to a lower utilization or heat dissipation temperature and results in a correspondingly higher thermal efficiency. However, higher compression ratios require higher working pressures, which requires more expensive motors. There is a point beyond which the unit cost of useful work quickly becomes excessive. In any case, the materials of manufacture of the engine limit the maximum pressure and the maximum temperature which pra se _{angewen (j e} can be t.

009887/1445 ~ ³ "

The charge volume of a known diesel engine is the volume of air which is contained in the cylinder when the piston is at bottom dead center, and the Volume headroom or the compression volume is the volume of air that is contained in the cylinder, if the piston is at top dead center. The ratio of charge volume to compression volume is called the compression ratio of the engine. The loading volume is equal to the sum of the compression volume and the suction volume. The suction volume is the volume sucked in by the piston as it moves from top to bottom dead center. During the compression stroke the air mass filling the charge volume is compressed to the compression volume. The· A decrease in volume causes both an increase in pressure and an increase in the temperature of the air. The relationship of final pressure and final temperature to initial pressure and initial temperature is exponentially proportional to Compression ratio. The final pressure is limited by the materials of the engine and is the initial pressure of course an atmosphere. The materials of the engine therefore limit the ratio of final to initial pressure and they also limit the compression ratio. Fuel is released into the air mass, which by compressing it has been heated to the compression volume, injected and burned with this air mass. The fuel is fed in at a controlled speed splashes so that the pressure is kept constant while the piston returns from top dead center. the Outward movement of the piston creates a larger volume in the cylinder, which is the original air mass must fill. To keep the air pressure constant at its maximum

009 8 8 7 / U

The burned fuel must maintain the air temperature increase accordingly to compensate for the increase in volume. The fuel is therefore at constant Pressure burned until the air temperature has reached a predetermined maximum allowable value. If the If the fuel supply is interrupted, the original air mass is at the same pressure that it decreases however, it has a larger volume and is much hotter than when compressing into the compression volume.

This hotter, larger volume air mass can only expand to the original cylinder charge volume. Therefore, the ratio of the feed volume to the volume at the fuel cut-off point is the expansion ratio, which is directly proportional to the feed volume and inversely proportional to the volume at the break-off point. The volume at the fuel break point is equal to the compression volume multiplied by the ratio of the break temperature to the temperature that was reached when the air was compressed into the compression volume. The expansion ratio is therefore much lower than the compression ratio. Since the pressure in the interrupted volume at the beginning of the expansion stroke is equal to the pressure in the compression volume at the end of the compression stroke ₉ , the gases remaining in the cylinder at the end of the expansion stroke are necessarily at a higher pressure and temperature than the air in the cylinder at the beginning of the compression stroke.

In known engines without pre-compression, these

00 98 87 / U4 5

Relatively hot exhaust gases under relatively high pressure are released directly into the atmosphere.

In known engines with pre-compression, the cylinder exhaust gases are used to drive a pre-compressor. This reduces the operating costs per unit of useful work output, since the part of the compression and expansion carried out with a large volume and low pressure is carried out in compact, cheap and very powerful centrifuge or turbine systems and the expensive piston system with low efficiency is only for the the high pressure, low volume part of the duty cycle is used. However, the theoretical thermal efficiency of the Motois apparently remains unchanged, since the pre-compression only increases the pressure level of the whole working cycle, and the pre-compressor theoretically emits its exhaust gases at the same temperature and pressure as a non-pre-compressed engine, which operates at the same peaks of temperature and Pressure Λ is working. The design of the motor still limits the peak values of pressure and temperature permissible in the duty cycle and consequently also the maximum theoretical thermal efficiency that can be achieved by the motor, regardless of whether a supercharger is used or not.

A theoretical thermal efficiency of about 55 % can be achieved with a modern diesel engine with a compression ratio of 15.6, which. Air with

- 6 009887 / U45

Ambient temperature and pressure are sucked off, the same compressed to Hl atmospheres and 660 ° C (1220 ° P), fuel to heat the air to 2504 ° C (45 ^ 0 ° P) burns and the gases from the cylinder at about k 1/2 atmospheres and ejects 1177 ° C (2150 ° P). It is possible to further expand the external pressure of these cylinder exhaust gases and thereby achieve a combined theoretical efficiency of about 60 % , but this must be done by driving power-consuming devices outside the working cycle «. "These external devices can make the improvement of the efficiency uneconomical, since in many cases the additional costs can not be justified by the saved fuel"

According to the invention, a four-day piston engine is used in which the cylinder compression volume is so is kept as low as possible. With the expression "minimal volume leeway" or "minimal compression volume" is the actual volume between the cylinder head and the piston at top dead center. This volume results from the purely mechanical clearance that is required between adjacent parts is. Unlike a known engine, the volume does not affect the compression ratio of the Motor, as it is not necessary that the load volume contained air is compressed to the compression volume. Every cylinder in the engine has a hot air and a cold air inlet valve and a hot air and a cold air outlet valve. In the case of a multi-cylinder engine, the engine is equipped with a hot air and a cold air intake manifold and a hot

009887714

air and a cold air outlet manifold. To compress the air, in addition to the Cylinders of the combustion engine a turbo supercharger, preferably with cooled suction air, which can be designed in several stages and as an additional compressor is working. Finally, there is an exhaust heat recuperator used, in which the fully compressed air preferably at least for the Auto-ignition of the fuel is heated to the required temperature. f

The first two strokes or strokes of the piston form the cooling portion of the duty cycle, using the cold inlet and cold outlet valves. It is noted that this means relatively cold, since in parts of these two cycles the temperature may actually be above ambient, but is much lower than the temperatures in the other two cycles in which the hot inlet and hot outlet valves are used will. However, for convenience of description, the terms "hot" and "cold" are used in this relative sense. The externally compressed "cold" μ air, which flows into the cylinder through the cold inlet valve during the first stroke of the operating cycle, contributes to the cooling of the internal parts of the engine, such as the piston and the cylinder wall. Additional cooling of the motor is of course provided by the usual water jacket or cooling fins.

The external exhaust gas heat recuperator is used for heating the fully compressed air preferably to a sufficient level:

- 8 009887 / UA5

high temperature triggering auto-ignition of the fuel. This embodiment of the invention which enabling constant pressure operation during the combustion portion of the duty cycle is one preferred embodiment and will be described first, but others will be pointed out below as well Embodiments are possible. A burn at constant Pressure is similar to what occurs in a typical diesel duty cycle, but it is note that the present invention is not concerned with a diesel duty cycle, although in this embodiment constant pressure combustion is applied. This will become apparent from the further description and also clear from a consideration of the greatly increased efficiencies made possible by the invention.

In a preferred embodiment of the invention Atmospheric air before entering the pre-compressor, d. H. the additional compressor, cooled. The pre-compressor exiting compressed air can advantageously be cooled again before entering the cold inlet manifold and from there it flows through the cold inlet valve into the cylinder. The operations of the exhaust, Inlet and fuel injectors are described here as being at the theoretically ideal point, for example at top or bottom dead center. A delay or introduction to the valve timing is usually desirable in practice.

The cold inlet valve opens at the top of the stroke and remains open while the piston first descends and then over the first predetermined portion

- 9 009887 / UA5

his return stroke comes back. As stated above, the cooled air is used to cool the inner parts of the engine including the cylinder and piston and as a result, they will be warmed up even though they are still well below ambient can. The air displaced by the piston during the first part of the return stroke is released into the cold inlet manifold returned through the still open cold inlet valve. The temperature level of the cooled air (( is set deep enough to compensate for any small heat input. At a predetermined point of the return stroke the valve is closed, and when the piston continues to increase, the trapped cool air remaining in the cylinder is further compressed. This included Air mass represents the amount of fluid actually working in the work cycle.,

When the air is compressed to the desired pressure, the cold outlet valve is opened at a second predetermined point of the return stroke. The trapped air in the cylinder can be located prior to compression on a Removing μ sufficiently low temperature so that it is located downstream of the compression ideally at ambient temperature. The piston continues to move until the end of its stroke, where the compression volume is very small and the high pressure air at ambient temperature is almost completely expelled from the cylinder. In one embodiment, the high pressure air is directed to the cold outlet manifold and flows through a heat exchanger recuperator in which it is heated to at least the temperature required for auto-ignition

- 10 OO9807 / U4S

- ίο -

is heated. In another embodiment, the high pressure air flows into the cold outlet manifold and is given in series by a low temperature ar ~ outlet heat recuperator and then through an outlet heat recuperator operating at high temperature directed. In this latter, working at high temperature heat recuperator, the temperature is on the Reduced value that an expansion or expansion machine, such as a gas turbine, can withstand, and the The recuperator operating at low temperature absorbs the exhaust gases from the turbine. The recuperatively heated High pressure air then flows into the hot inlet manifold.

At the top of the stroke, the cold outlet valve is closed and the hot inlet valve is opened and When the piston descends, the recuperatively heated air flows from the hot inlet manifold into the Cylinder.

Fuel is injected at a controlled rate along with the heated air and burns when injecting. This gradual injection and burning of the fuel can be used in known engines cannot be achieved. In known engines, the fuel is injected into the entire working air mass and the result is an explosion and incomplete combustion. In a preferred embodiment the fuel burns at essentially constant pressure and increasing temperature, while the piston lowers. At a predetermined spark in the power stroke or power stroke, the hot inlet valve is

- 11 -

closed, the fuel supply interrupted and the Gases continue to expand isentropically, being both Losing both temperature and pressure while using the Push the piston to the bottom of its stroke. The hot outlet valve is opened at about the point on which the piston begins its return stroke, whereby the piston at the end of the stroke almost completely the exhaust gases from the cylinder into the hot outlet manifold can eject.

In one embodiment of the invention, these flow hot exhaust gases through an expansion or relaxation turbine, which drives the pre-compressor, which acts as an auxiliary compressor is used. In this embodiment, the amount of fuel injected must be so controlled ensure that the temperature of the exhaust gases does not exceed the temperature that the turbine can withstand. The exhaust gases then flow through an exhaust heat recuperator, from which they are approximately at ambient pressure and temperature leak to atmosphere.

In a further embodiment of the invention flow the hot exhaust gases first through a recuperator working at high temperature, in which their temperature is lowered to that which the turbine can withstand can, and then they are through the expansion turbine which drives the pre-compressor. The gases from the turbine flow through one at lower Temperature working exhaust heat recuperator, from which it is approximately at ambient temperature and pressure to the atmosphere as in the case of the

- 12 -

benen embodiment of the invention. The relatively cool, high-pressure air flows
first by the one working at low temperature
Recuperator and then in series through the high temperature recuperator.

Since the engine works with a very low compression volume, the exhaust gases are almost at the end of the stroke
completely ejected from the cylinder and the. Duty cycle can be repeated. It is noted that the
hot exhaust gases useful and practically all of their heat
release reversibly, either in the form of work in the expansion turbine driving the pre-compressor or in countercurrent in the recuperators by warming up relatively cold, compressed air on the
temperature required for auto-ignition. This is one of the important features of the present invention since a large amount of heat is lost in the exhaust gases in an ordinary diesel duty cycle. In the working cycle according to the invention, however, this heat is in
effective use of a simple system.

The second embodiment, in which a high and Using a low temperature recuperator will significantly increase the power output of the Motor according to the invention with a slight decrease in thermal efficiency, which is desirable for many purposes.

Preferred exemplary embodiments have been described above, but other embodiments are also possible by adding a certain. Degree of profitability or

- 13 009887 / 1U5

203S605

.- 13 -

Performance is sacrificed, but results are still achieved that are much better than can be achieved with normal work cycles. For example, all of the air compression can be done outside of the reciprocating engine, but this increases the cost. It is also possible to carry out all of the compression by the piston engine, which also means an increase in costs. Instead of a turbo compressor that works with cooled suction and air, multi-stage similar can be used

external compressors with intake of air at ambient temperature and with intermediate cooling to ambient temperature can be used, again the cost increases slightly and the heat efficiency slightly increases sinks. It is also possible to cool the suction air of the turbo compressor to such a low value that without further cooling the suction air temperature of the reciprocating compressor has the desired low value. It is not essential that the combustion takes place at constant pressure, with auto-ignition takes place, however, since the constant volume combustion has the same high efficiency generated at lower useful work output. It is also the use of a lower recuperator temperature in connection with spark plug or glow plug ignition possible, but this leads to a loss in thermal efficiency.

The invention is illustrated by way of example using the figures explained in more detail. It shows

FIG. 1 shows a schematic illustration with a section through a cylinder of an internal combustion engine according to FIG Invention, the associated parts are also shown schematically,

009887/1445 - ik -

203S605

Figure 2 is a working diagram of the engine shown in Figure 1 with the different piston positions,

Figure 3 is a valve diagram for an embodiment of the Invention,

Figure ^ a temperature-entropy diagram of a diesel work s cycle,

FIG. 5 is a similar diagram of an embodiment of FIG Invention in which multi-stage, inter-cooled, external compressors are used,

Figure 6 is a similar diagram of a preferred embodiment of the invention, in which combustion at constant pressure is applied,

FIG. 7 shows a similar diagram of a further embodiment of the invention, in which combustion at constant Volume is applied

Figure 8 is a similar diagram of an embodiment a normal diesel duty cycle and

FIG. 9 shows a modification of the embodiment shown in FIG. 5, showing the recuperators operating at high and low temperatures.

Figure 1 shows a schematic section through a cylinder of an internal combustion engine according to the invention. The cylinder 1 with a conventional water jacket 26 has a piston 2, which over part of its way

- 15 009887/1445

is shown shifted downwards on the intake stroke for the first or cold intermediate compressed air. The rotation of the crankshaft takes place in a clockwise direction, as can be seen from FIG. A series of four poppet valves 6, 7, 8 and 9 are provided, the valve 6 serving for the intake of cold intermediate compressed air and the valve 7 serving for the intake of hot high-pressure air. The valve 8 is an exhaust valve for hot, is under relatively low pressure exhaust gases and the valve 9 is a similar exhaust valve for cold high-pressure air. A fuel injector 10 is provided and the four valves are connected to manifolds 16, 17, 18 and 19, respectively. Although only one cylinder is shown in the figure, in reality the internal combustion engine usually has several cylinders and therefore distribution lines are shown. Since Figure 1 shows the piston near the beginning of its stroke to the cold air inlet, only valve 6 is open. In general, the valves are conventionally designed and the actuating devices, such as cams, camshafts and the like, are not shown since their design is not influenced by the invention. Of course, the cams must operate the ™ valves according to the work cycle to be described.

Figure 2 is purely schematic and shows the position of the upper end of the piston at a number of points in the working cycle. To make the figure clearer, the schematic representation of only the upper end of the piston. In Figure 2, the four strokes are shown in Roman numerals. set, namely I for the inlet of cold intermediate compressed air, II for the outlet of cold intermediate pressure gas

. - 16 009887 / 1U5 .

and then high pressure compression and discharge, III for the inlet of hot high pressure air, the combustion of the fuel and subsequent isentropic expansion, and IV for the discharge of the hot, relatively low pressure exhaust gas. The stroke I runs from the top dead center, where the margin is very small, to the bottom dead center. The top dead center is with (T), the bottom dead center with CD »intermediate positions in the upstroke for II with (2a) and (2b), the top dead center with (3), an intermediate position in the power stroke III with Q>&) , the bottom dead center is denoted by (] y and finally the upward boom IV runs from bottom dead center to top dead center (X) , and this scope also shows the top and bottom dead center position and the intermediate positions for the various strokes of the four-stroke cycle.

Figure 3 is a valve diagram is self-explanatory mk, wherein the piston positions of the top and valve numbers are indicated along with their labels on the left side, and wherein the two lines each with. O and C, respectively, correspond to open and closed periods. The diagram also shows the opening and closing periods for the fuel injector 10.

At the beginning of the stroke I, which is shown in Figure 1, atmospheric air is through the cooler 25 under the

- 17 0 0 9 8 8 7 / U 4 5

Suction of the turbo compressor 22 cooled, which of the Shaft 21 is driven as described below. The emerging through line 23 compressed Air is once again brought to a temperature by the cooler 24 cooled below ambient temperature. The cold intermediate compressed air is not on the highest pressure of the duty cycle, since, as described below, the final compression by the piston 2 takes place in cylinder 1 of the engine. The coolers 24 and 25 increase the compression efficiency. In addition, the temperature of those leaving the cooler 24 lies Cold air lower than the ambient temperature, so that the inner parts of the cylinder are more strongly cooled, than they would be cooled by ambient air. There the maximum allowable operating temperature is the mean of the peak temperature and the inlet temperature a decrease in inlet temperature a corresponding one Increase in the peak temperature. The cooler can be designed in the usual way and therefore no details are shown. They are pure in Figure 1 shown schematically as rectangles. You can also be designed in several stages, so that they approximate a reversible operation.

When the piston lowers during stroke I, cold intermediate compressed air flows into the cylinder through valve 6. The piston performs a full stroke down to the bottom. ren dead center © and then rises to the intermediate position (| a) during stroke II. From Figure 3 it can be seen that the valve 6 was open during this entire time. Between the positions (g) and (gjp there is a

- 18 009 8 87 / U45

Part of the cold intermediate compressed air into the distribution line 16 blown back. This is a preferred embodiment because it gives maximum cooling.

When the piston reaches the point (§a) shown in Figure 2 the valve 6 is closed, whereby the remaining air is trapped in the cylinder. The point (^ a) is determined by the air mass, which in the following Power stroke is to be processed. It is noted that the intermediate positions (5a), (2b) and Qa) are not are the same. If the piston continues to rise from point (2a) to point (2b) during stroke II, then the enclosed, relatively cold intermediate compressed air is further compressed up to the peak pressure of the working cycle. At this point, their temperature has usually reached ambient temperature. The valve 9 is then opened and when the piston continues its movement, the high-pressure air at ambient temperature is almost completely out of the cylinder into the cold one Outlet distribution line 19 and from there through line 15 into an exhaust gas heat recuperator 12, wherein the partially heated gases are then passed through the pipe 29 to a high-temperature recuperator 27, where they are heated to the final temperature. These recuperators are normal heat exchangers, e.g. pipes, around which the hot gases flow, and they are therefore only indicated schematically in FIG. The recuperator ' Ren heat the high pressure air, which is at ambient temperature, to a temperature which is sufficient is high so that auto-ignition of the fuel can occur. The hot high pressure air leaves the recuperation

- 19 -

009087/1445

rator 27 and enters the distribution line 17 through the line 20 a.

When the piston begins to move down from top dead center position (D during stroke III, the valve 7 is opened, as can be seen from Figure 3, and lets in the hot air. At the same time will Fuel is injected in a controlled amount through the injector 10, as shown in Figure 3, and burns at constant pressure while the piston continues to return to position (3a). The fuel injector 10 is then closed, valve 7 is closed and the remaining power stroke III is set theoretically represents an isentropic expansion.

When the piston reaches bottom dead center (T) and begins to rise in its exhaust stroke IV, the valve 8 is opened for the hot exhaust gases. The valve 8 is shown in FIG 3 referred to as the outlet valve for relatively low pressure. The gases are at a very high temperature and the pressure is still quite high, e.g. a little below 2 atmospheres, but is in comparison at the peak pressure of the duty cycle, which prevailed at the beginning of stroke III, low. The flow from valve 8 hot gases into the exhaust manifold 18 and into one High-temperature recuperator 27 and from there through the pipe 28 into the turbo-expansion valve 11. The high-temperature · Recuperator 27 lowers the temperature of the exhaust gases sufficiently so that the turbo-expansion valve 11 can process them can. This relaxer operates the shaft 21, which according to Figure 1, the turbo compressor 22 drives. In the

- 20 0098-87 / UA 5.

Turbine 11, the hot gases relax to approximately atmospheric pressure and enter recuperator 12 through line 13. They are still quite hot and are used to heat the incoming cold high pressure air in countercurrent, as explained below. During this preheating, the temperature of the hot gases drops to slightly above ambient temperature and the gases exit through the exhaust pipe I ¹ I to the atmosphere. The highest temperatures encountered in the duty cycle are generally determined by the maximum temperature that the materials in the various parts can withstand and are chosen so that each part of the plant is operated under safe, but near optimal, conditions. It is noted that the high-temperature recuperator 27, which contains no moving parts, can withstand much higher temperatures than the rapidly rotating turbine 11. This enables a higher end temperature in the working cycle and, as stated above, the overall efficiency with only a slight decrease Noticeably increased performance.

At this point it should be noted that the Duty cycle represents a major improvement over the normal diesel duty cycle, since the Exhaust gases are approximately at ambient temperature and pressure. It also causes the use of refrigerated Air during stroke I has a very strong cooling of the inner parts of the cylinder. This allows the power swing with a much higher peak temperature than would be possible if the pistons and cylinder walls if they were cooled by much warmer air, they would be hotter, as is the case with the well-known diesel work

- 21 009887 / 1A45

cycle the case, 1st. It has been explained above that the theoretical efficiency of the duty cycle of the maximum permissible peak temperature and the achievable depends on the minimum exhaust gas temperature. In both respects, the duty cycle of the present invention is significantly better than even the best diesel duty cycle.

The advantages of the present invention can be demonstrated by Figure 4, which is a temperature-entropy diagram shows a normal, efficient and modern diesel duty cycle, FIG. 8, which shows a similar one Diagram of the same duty cycle using the cylinder exhaust energy reproduces and the figure 6, which is a similar diagram of the preferred embodiment of FIG Invention reproduces. In each diagram, the amount of useful work generated is proportional to the enclosed Duty cycle area. In Figures 4 to 8 these are Areas lightly hatched to express their differences more clearly.

In Figure ^1, the duty cycle 1st J of a normal, modern diesel engine having a compression ratio of 15> represented by the curve 6 AHJKA. Point A corresponds to 1 atmosphere pressure and approximately 38 ° C (100 ° F), H corresponds to 47 atmospheres and 6k9 ° C (1200 ° F), J corresponds to 47 atmospheres and \ 2504 ° C (4540 ° F) and K. corresponds to 4.66 atmospheres and 1177 ° C (2150 ° P). The path AH corresponds to an isentropic compression, HJ to an isobaric heat increase, JK to an isentropic expansion and KA to an isohoric

- 22 009887 / U45

Heat dissipation. The theoretical efficiency is about 55 %. It is noted that theoretical efficiencies are used when comparing the duty cycles. The efficiency of a real engine is less than theoretical due to the pressure drop, friction and other mechanical losses in the engine.

In Figure 8, point K corresponds to a temperature of 1177 ° C (2150 ° F) and a pressure of 4.66 atmospheres. Theoretically, this air can afford an additional work when it is mixed or passed through a turbo expander, but in practice it is cooled too hot for use and must first isochoric to about 8L6 ° C (1500 ° F), and 2 _a 27 atmospheres before it can be isentropically expanded to one atmosphere and 1 »93 ° C (920 ° F) in a turbo-vent. In Figure 8, this isochoric cooling is the path between K and M and the isentropic expansion is the path between M and N. The cycle is represented by the line AHJKMNA, with points A, H, J and K being the same as in Figure 4. The compressor exit air temperature at point H is about 167 ° C (300 ° F) higher than the turbine outlet temperature at point N, making it impossible to recover exhaust heat and this is one of the major inevitable disadvantages Diesel work cycle. The use of an expansion turbine from point M to point N theoretically increases the efficiency to about $ 60, but only if the work energy extracted from the cylinder exhaust gases is used to drive a generator or other work. needing device outside of the fle® Itotomrbelts cycle

- 23 7 / U45

is used. When the work energy to drive
a supercharger is used for the engine, theoretically no increase in efficiency can be achieved.

The diagram AGHJKLA in Figure 6 is the duty cycle of a preferred embodiment of the invention with a theoretical thermal efficiency of about 82% again, which is almost 50% higher than the ones shown, presented in FIG M ä duty cycle of a normal diesel engine. Point A corresponds to one atmosphere and 38 ° C (100 ° F) as in Figure 4. Point G after compression corresponds to kl atmospheres and 38 ° C (100 ° F). The preheating in
outer recuperator 12 brings the air at point H to kj atmospheres and about 6 ^ 9 ° C (1200 ° F). During the
In the fuel injection section of the power stroke, the temperature rises to 2504 ° C (Λ5 * »0 ^ο F) at point J. Es
it is noted that this is the same temperature that was reached in the normal diesel duty cycle in FIG. The temperature was standardized, to facilitate comparisons of the various working cycles and it was drawn, no advantage from the fact that the piston and m cylinder walls additional cooling by the cold high-pressure air during the strokes I and receiving II and that according to the invention, the point J could be at a higher temperature during safe operation. The point 0
corresponds to 1.78 atmospheres and 816 ° C (1500 ° F) and "the point L corresponds to one atmosphere and 6M9 ⁰ C
(1200 ° F). The way AG corresponds to an isothermal compression, which can be closely approximated by using a sufficient number of cooling stages. The compression can

- 24- -

009887 / U45

take place partially or completely outside of the working cylinder. The path GH corresponds to a recuperative isobar heat increase. HJ corresponds to a direct isobaric heat increase through combustion of fuel, as in the case of FIG. 4. JO corresponds to isotropic expansion in the engine cylinder, 0-L corresponds to isentropic expansion in the turbo-expansion valve and LA corresponds to recuperative isobaric heat dissipation. The air compression AG can be carried out entirely outside the reciprocating engine, but the compression section with the highest pressure increase is most economically carried out in the engine cylinder itself. If the compression is carried out with multi-stage cooled suction air, the exhaust gases of the engine cylinder can provide energy to operate the supercharger so that it achieves a compression ratio of about 6.6. Therefore, of the total required compression ratio of 47.0, the cylinder only needs to deliver a ratio of 7.1.

The heating of the air in the recuperator at constant pressure causes a significant increase in the air volume. This increase in volume is done in order to work too by letting the air flow in through the hot air inlet valve and moving the piston of the engine. In the case of a multi-cylinder engine with distributor lines, which is preferred, the one withdrawn from the distributor line Amount of air replaced at the same time by a supply of fresh air, which in one of the other cylinders has been compressed and heated in the recuperator. The pressure in the distribution line therefore remains almost constant.

- 25 009887 / UA5

The temperature at point J on the graph is the peak temperature of the duty cycle. The pressure at the end of the piston stroke, which is shown at O, is approximately 1.78 atmospheres and the temperature is close to 8.16 ° C (1500 ° F). The minimum volume headroom or the compaction volume, which is a feature of the present Invention forms, enables the piston to close the high temperature exhaust gases from the cylinder pushes out completely.

According to the invention, the expansion ratio is as in known engines the ratio of the charge volume to the volume in the cylinder when the fuel supply is interrupted contained air. According to the invention, the expansion ratio is completely independent of the compression ratio or the maximum permissible operating pressure. Rather, it is determined by the volume of the under high pressure heated air which is in the cylinder is included when the fuel supply is interrupted. This volume can be chosen so that gives any desired cylinder outlet temperature. The cylinder outlet temperature is the same as the inlet temperature of the turbo-expansion valve and in normal operation this temperature is limited to a maximum of about 8l6 ° C (1500 ° P) limited. In known motors, the required temperature reduction is a waste of energy Blow off is achieved and the cylinder exhaust valves are. also subjected to a much higher temperature.

Theoretically, the only thermal irreversibility in the preferred embodiment of the invention is the difference between the temperature of 6 ^ 9 ° C (1200 ° P)

- 26 - 00988 7 / 1U5

the high pressure air leaving the recuperator (point H) and the split temperature of the working cycle of 2504 ° C (4540 ° P, point J). This irreversibility limits the theoretical thermal efficiency to 82 % increase to about 89 % , which corresponds exactly to the efficiency of a Carnot cycle carried out between the same temperature limits of 2504 ° and 38 ° C (45 ¹ IO ⁰ and 100 ° P). However, in order to achieve maximum efficiency, it is still important that the temperature of the air leaving the recuperator is as high as possible. The usual recuperator training limits this temperature to about 649 ° C (1200 ° P). The turbo ventilator must therefore operate with an inlet temperature of 816 ° C (1500 ° F) and an outlet temperature of 649 ° C (1200 ° P). This theoretically requires isentropic expansion from 1.78 atmospheres pressure to one atmosphere pressure.

A minimum volume clearance was defined above as the actual volume between the piston and the cylinder head, which is based on the purely mechanical clearances required between these parts. The cold air remaining in this volume at the end of the compression stroke cannot be heated in the recuperator. This reduces the efficiency of the duty cycle. During the following power stroke, this cold air mixes with the preheated hot air and cools

- 27 009887 / U4S

same. This also reduces the efficiency of the duty cycle. Those in this volume on, ' The hot exhaust gases remaining at the end of the exhaust stroke mix with the incoming cooled air and warm this on and therefore the heat of the exhaust gases is not available in the recuperator. This further reduces the efficiency of the duty cycle. The incoming Air must be cooled to a lower value in order to compensate for this undesired heat input, and this again increases the efficiency of the duty cycle decreased. In order to achieve maximum efficiency, it is therefore desirable, but not essential for operation, for this minimum volume margin to be so small as possible.

The maximum temperature at point J is also limited, in this case by the design of the piston engine. The optimal thermal efficiency results when an isentropic, labor-producing temperature drop of the maximum permissible peak temperature of the working cycle (J) to the maximum permissible recuperator temperature (L) is present. For the given temperature conditions Theoretically, this temperature drop occurs when the initial pressure of the entire expansion process is the 47 times the outlet pressure. This is the optimal one Condition. If a higher initial pressure is used, the temperature at point L will exceed the maximum ' permissible value, which requires isochoric cooling, and the thermal efficiency in turn suffers underneath.

The presence of an optimal initial expansion

- 28 0 0 9 8 8 7 / U 4 5.

pressure, which is also the optimum peak operating pressure, is only one of the present invention. coming feature. Although the initially low thermal efficiency of known engines increases with each increase the operating pressure is increasingly improved, would be an impossible to meet operating condition of more than 2000 atmospheres at 2504 ° C (4540 ° P) required, for a known engine to approximate the thermal efficiency of the preferred duty cycle according to the invention.

In another embodiment of the invention, the same theoretical thermal efficiency of 82 % is achieved when the fuel is burned at constant volume rather than constant pressure. In this constant-volume combustion, cooled suction air is also compressed and full countercurrent preheating takes place at the same maximum temperature and the same maximum pressure as in the previously described preferred embodiment of the invention. Instead of the maximum mechanical compression of 47 atmospheres required in the constant pressure process, the constant volume process requires mechanical compression of only 15.8 atmospheres. The remainder of the final pressure of 47 atmospheres results from the supply of heat at constant volume to the air contained in the cylinder. The recuperator must therefore be designed so that it can withstand a pressure of only 15.8 atmospheres instead of 47 atmospheres and therefore the recuperator, the distribution lines, valves and pipes are much cheaper. However, the useful work power of an engine with a given suction power is slightly less than in

- 29 -

(109887 / U4 5

the preferred case and the choice of constant pressure or constant volume method can be based on based on a macroeconomic assessment. The heat supply at constant volume is different away from the heat supply at constant Pressure in the timing and speed of fuel injection. When using constant By volume, the fuel is injected quickly, theoretically instantaneously, after the piston moves has moved over part of the power stroke and after the hot air inlet valve has closed or around the time of this closing. When using At a constant pressure, the fuel is injected simultaneously with the preheated combustion air at a regulated speed in order to keep the pressure constant keep.

This embodiment of the heat supply at constant Volume is shown with the curve A-C-E-G-A in Figure 7, A-G corresponds to an isotherm, C-D to an isobar, D-E an isochore, E-G an isentropic and G-A an isobar. As above, point A corresponds to an atmosphere and 38 ° C (100 ° F), point B corresponds to 6.6 atmospheres and 38 ° C (100 ° F), point C corresponds to 15.8 Atmospheres and 38 ° C (100 ° F), point D corresponds to 15.8 atmospheres and 649 ° C (1200 ° F), point E corresponds to 47 atmospheres and 2504 ° C (4540 ° F), the point F equals 1.78 atmospheres and 816 ° C (1500 ° F) and point G corresponds to one atmosphere and 649 ° C (1200 ° F). The way A-B corresponds to the isothermal work, which are required for the external turbo compression and

- 30 -

0.3 '.; :: 7 / I 4 45

equal to that of the external isentropic turbo-relaxation is derived work. The path B-C corresponds to the required isothermal compression work of the piston engine. Path C-D corresponds to the external isobars of recuperative heat input. The path D-E corresponds to direct heat supply at constant volume, the Path E-P corresponds to the isentropic expansion of the piston engine, path P-G corresponds to the outer isentropic Turbo relaxation and the path Q-A corresponds to the isobars of recuperative heat dissipation.

In a further embodiment of the invention, multistage compressors which are mutually cooled to ambient temperature are used, as shown in the curve ABCDEP in FIG. The theoretical thermal efficiency of this work cycle is about 76%, the paths AB and CD correspond to the external compression to a total of 5.3 atmospheres and the path EP corresponds to the engine cylinder compression from 5.3 to 47 atmospheres, the ones for the external additional compression steps AB and CD The energy required is generated by the final expansion of the engine cylinder exhaust from point O (1.78 atmospheres, 81.6 ° C or 1500 ° P) to point L (1 atmosphere, 649 ° C or 1200 ° F) in a suitable high-performance turbo system delivered. Step EP, which corresponds to a compression ratio of approximately 8.9, must be carried out in the engine cylinder. (In the preferred embodiment of the invention, the compression work is reduced by the use of cooled suction air, so that the same amount of work energy, which is recovered from the cylinder exhausts, in the preferred

- 31 0 0 9 8 0 7 /! kh 5

This embodiment gives a greater external compression ratio than this embodiment). the Turbine outlet temperature (point L) is about 355 ° Q (640 ° P) higher than the compressor outlet temperature (point P), allowing a significant amount of exhaust heat recovery is possible.

The theoretical efficiency of 76 % is not quite as high as in the preferred embodiment described above and shown in FIG. 6, but it is still much higher than that which can be achieved by the best known motors. The omission of cooling may in certain cases be of economic interest, and it is an advantage of the present invention that various embodiments can be used in all cases to achieve the best economic compromise.

In FIG. 9, the curve AFCGA corresponds to a normal diesel work cycle and the curve ABCEA corresponds to a work cycle according to the invention, with BF corresponding to a recuperator operation at low temperature. Lines of constant volume are also shown in this diagram and it can be seen that point D ^{lies on the line of 22,654 1 (800 ft 3} ) volume. From the above description it can be seen that point D also corresponds to the suction power of the engine. Point G is on the 11,554 1 (408 ft ³ ) volume line for the normal diesel duty cycle. An engine designed according to the invention therefore requires almost twice the suction power of a normal diesel engine per kilogram mole of processed air and although it has one per kilogram mole

- 32 - 00.9887 / UAS-.

does much more work with much greater efficiency, its mean effective pressure (BMEP) is only 80 % of that of a diesel engine.

The curve ABKLCGHJA corresponds to another embodiment of the invention, where BKJA is a recuperative effect at low temperature and KLGH is a recuperative effect at high temperature. The theoretical thermal efficiency of this modified duty cycle is about 80.5 % and its mean effective pressure is about 40% higher than that of normal diesel duty cycles, which have a thermal efficiency of only 55 % . In this embodiment of the invention, point L is theoretically about 771 ° C (1 ^ 20 ° P), which is higher than is permissible with currently available recuperators. In practice, however, there must be a temperature difference between the exhaust gases and the incoming compressed air, so that in practice the temperature of the incoming compressed air does not exceed the somewhat flexible limit of about 619 ° C (1200 ° P).

- 33 009887 / U45

Claims (8)

Claims ^% / 1] Four-stroke internal combustion piston engine with a fuel v- <_ feeding means, characterized by means for compressing air outside of the motor and for cooling the compressed air to a predetermined temperature, two sets of intake and exhaust valves for each cylinder including an intake valve for admitting cooled compressed air into the cylinder during the intake stroke of the four-stroke duty cycle of the engine, whereby the relatively cold intake air helps to cool the engine cylinder, the intake valve opens near top dead center of the intake stroke and closes at a point at which a predetermined Air working volume is reached, a cylinder outlet device for the cooled air, which is closed during the entire suction stroke and open during the compression stroke when the compressed air is at the peak pressure of the working cycle, a heat exchanger device for two gas flows, which the gases röme holding separated from each other, wherein the heat exchange means having an inlet means for the cooled compressed air from the cylinder under the peak pressure of the working cycle and an outlet for heated compressed air and the compressed air flows through the heat exchanger means as one of the two gas streams, whereby the cooled, compressed air is heated becomes, a second cylinder intake valve - 3 * OQ9887 / 1U0 for introducing the heated compressed air from the heat exchanger means, wherein the second inlet valve opens approximately at top dead center of the next following power stroke, so that in the heat exchanger device heated compressed air is introduced into the cylinder, and the second inlet valve is opened remains until a predetermined volume of hot compressed air has entered the cylinder, and then is closed, means for introducing fuel and burning it in the cylinder a second during at least a portion of the period during which the second intake valve is open Cylinder valve which opens near the bottom dead center of the power stroke and during the next one Exhaust stroke remains open, an external gas expansion machine, such as a gas turbine, and a device for Introducing hot exhaust gas, which arises in the cylinder during the power stroke, so that the hot ones are below High pressure exhaust gases in the machine relax to around ambient pressure and generate power, as well by a work-performing device for connecting the outlet of the expansion machine to at least one Part of the heat exchange device, so that the under low Pressurized exhaust gases from the machine are cooled. den and their heat to those at the peak pressure of the working cycle release any chilled air. 2. Internal combustion engine according to claim 1, characterized in that that the heat exchange device has two parts through which the one located at the peak pressure of the working cycle cooled high pressure air flows in series and is heated, with the hot exhaust gases coming out of the wide exhaust valve - 35 - 009887 / 1US - 35 - ■. ■■■■ * flow through the first of the two parts of the heat exchange device, this heat exchange device Has sufficient dimensions to lower the temperature of the hot exhaust gases so that the expansion machine can withstand the same, and wherein the second part with the outlet of the expansion machine is connected so that the flow of the cooled high pressure air at the peak pressure of the duty cycle flows in series first through the second part and then through the first part. 3. Internal combustion engine according to claim 1 and 2, characterized in that it is a multi-cylinder engine and cold air inlet and outlet distribution lines and hot air inlet and hot exhaust manifolds are provided which are connected to the cold air and hot air inlet and outlet Exhaust valves are connected so that shock waves in the flow of hot and cold gases are as low as possible. k. Internal combustion engine according to Claims 1 to 3, characterized in that the distance between the piston and the cylinder head at the top dead center is not significantly greater than the minimum clearance required to avoid mechanical contact between the adjacent parts. . · 5. Internal combustion engine according to claim 1 to 4, characterized in that that the gas expansion machine drives the external air compression device and at least one Provides part of the performance for the compression. 6. Internal combustion engine according to claim 1 to 5, characterized -36-009887 / 1U5 203SSQ shows that the compressed air leaving the heat exchange device is at a sufficiently high level Temperature is to cause auto-ignition when the fuel is introduced. 7. Internal combustion engine according to claim 1 to 6, characterized in that that the introduction of fuel begins approximately at the point in time at which the second inlet valve is opened and that the combustion during fuel injection is essentially under constant pressure takes place. 8. Internal combustion engine according to claim 1 to 6, characterized in that the fuel introducing device introduces fuel after a predetermined volume of hot compressed air has entered the cylinder through the second inlet valve, and that the fuel is burned at such a rate that the pressure in the cylinder increases. 009887 / 1U5 Blank page