Patent Application
20240044284
We on earth (8.0 billion people in 2022) are destroying our planet by wasteful use of resources, and in particular by wasting energy. Of all the vast production of energy on the planet, now at over 600×1015 BTUs (630 EJ) per year, about 80% is from burning fossil fuels at very low efficiency. Renewable energy sources are beginning to replace fossil fuel use, but the best solution in the near term, to meet our energy needs with far less dependence on fossil fuels, is to improve energy efficiency of fuel use. Our invention—a high efficiency engine—is directed to that purpose. It is particularly useful for combined heat and power applications where a combined efficiency of 90% or more may be achieved. Our engine, or engines, are ideally suited for use with renewable fuels.
The scientific principles of operation of internal combustion engines have been known for approximately 130 years, after Rudolph Diesel first applied the concept of the thermodynamic cycle in 1892, just 16 years after the foundation concepts were introduced by Willard Gibbs. Modern theory of the thermodynamic cycles of internal combustion engines began with Diesel's work. In Diesel's U.S. Pat. No. 608,845, he presents what has become known as the “Diesel cycle.” Today, the five well-known internal-combustion engine cycles are represented by standard reversible forms composed of isentropic, isochoric, and isobaric process steps. Those five cycles are: Diesel cycle, Otto cycle, dual cycle, Brayton cycle, and the Atkinson (or Miller) cycle. It was not generally known until recently that a sixth comprehensive standard thermodynamic cycle includes and extends the five prior cycles—we refer to this improved cycle as The General Cycle.
A thorough description of the General Cycle is provided in the reference: Ernest Rogers, “Calculating Engine Efficiency with the General Cycle Equation,” May, 2020, available on-line at the following web address: https://www.researchgate.net/publication/341133935_Calculating_Engine_Efficiency_with_the_General_Cycle_Equation
The above referenced paper by one of the applicants is reproduced substantially in its entirety herein.
Calculating Engine Efficiency with the General Cycle Equation
Ernest Rogers ⋅May 4, 2020
As used here, a thermodynamic cycle is a sequence of changes in the conditions of a gas; the final step returns the gas to its initial condition. The General Cycle will be described in terms of reversible changes of an ideal gas. Most heat engines can be analyzed by use of such an “ideal” cycle. Only the essential steps are included—for example, in representing a four-stroke engine, the two strokes for exchanging the gas will be ignored.
The General Cycle has this name because it is capable of representing most commonly-used internal combustion engines, such as carbureted gasoline engines, Atkinson engines, diesels, and even gas turbine engines. By assuming that gas properties—the specific heats and specific heat ratio—are constants, a very simple formula can be obtained for heat engine efficiency. This simple formula is remarkably accurate in predicting the efficiency of real engines when 1.35 is used for the “constant” value of the specific heat ratio, and an energy loss factor is judiciously applied. (The analysis leading to the formula is for a reversible cycle with no heat or friction losses.)
One may ask, why is this formula for efficiency needed? The answer is that it is a teaching tool that shows us how to develop more efficient engines.
The steps of the cycle are shown on the P-V Diagram (
I. Starting at point 1, a gas is compressed adiabatically (without heat transfer) from V1 to V2. The compression ratio is RC=V1/V2. The pressure increases from P1 to P2. The compression work from point 1 to point 2, W12, is negative.
II. A first heat (fuel) input Q1 raises pressure from P2 to P3 at constant volume. This P3 is the maximum pressure. No work is done and V3=V2.
III. A second heat input Q2 is added at constant pressure as the piston begins to move outward from V3 to V4. (Fuel began to burn at point 2 and burning is complete at point 4.) The total heat input is QIN=Q1+Q2. The work from point 3 to point 4 is W34.
IV. The gas expands adiabatically from point 4 to point 5. The power stroke ends at point 5. The expansion ratio RE=V5/V2 exceeds the compression ratio by the factor A=V5/V1. A is the Atkinson ratio. The work in this step, from point 4 to point 5, is W45.
V. Heat is removed at constant volume. (V6=V5) Pressure decreases from P5 to P1, the initial pressure. (P6=P1)
VI. The gas is compressed and heat is removed at constant pressure. The volume decreases from V5to V1, the initial volume, and the temperature returns to the initial temperature, T1. (V6=V5) The work from point 6 to point 1, W61, is negative.
The cycle is complete. The total heat removed in steps V and VI is the rejected heat, QOUT. The total work available from the cycle is W=W12 +W34 +W45 +W61. In the ideal cycle, W=QIN−QOUT. In_a real engine, the process is a little different than this ideal cycle—the steps will not be so neatly defined. The real engine is expected to have valves; for example: valves open at point 5 to remove exhaust gas. A fresh charge of air enters and the piston returns to point 1, the starting point. Then the valves are closed and a new cycle begins. Opening of valves in part of the cycle can cause a loss of work, as work against the atmosphere.
The efficiency of the cycle is obtained by comparing the total work W to the total heat input QIN. Efficiency is a dimensionless quantity. The efficiency of this ideal cycle can be expressed in terms of a set of defined dimensionless parameters for the cycle. The equation for cycle efficiency is:
Where η is the ideal efficiency,
As already mentioned, this equation encompasses most common engine cycles. If β=1, the equation simplifies to the equation for the Atkinson cycle. If this is further restricted to A=1, it becomes the Otto cycle. By setting A=1 only, you obtain the dual cycle. If you set α=1 and A=1, you obtain the classical Diesel cycle. If you set α=1 and A=β it becomes the Brayton cycle.
In order to make good use of the General Cycle equation, one must have some understanding of what limits should be placed on the selectable parameters, α, β, A, and RC. Then the equation can be the starting point of a search for a more efficient engine.
In order to describe our invention, it will be necessary to review the scientific principles pertaining to it and to define terms. As currently practiced, our invention is a two-stroke direct-injected piston engine that is represented by the General Cycle. Nevertheless, the General Cycle is also applicable to four-stroke engines. In a four-stroke engine, the two strokes that transfer the exhaust out of the engine and intake a fresh charge of air may be ignored, while the other two strokes, the compression and power strokes, are represented in the General Cycle. General Cycle
The General Cycle is an idealized thermodynamic cycle that can represent most, if not all, common internal combustion engines. Usually it is analyzed as a sequence of reversible steps performed on a compressible working fluid. In a real engine, this compressible fluid is a gas comprising oxygen with any amount of other gases, such as air or a gas composed of air, fuel, or combustion products. The General Cycle is best understood by reference to the P-V diagram of
I. Starting at point 1, a fresh charge of compressible fluid is compressed from volume V1 to volume V2. The compression ratio is RC=V1/V2. Pressure increases from P1 to P2 . The compression work from point 1 to point 2, defined as W12, is negative.
II. Beginning at point 2, a first heat input Q1 from fuel raises the pressure from P2 to P3, at constant volume. This P3 is the maximum pressure, and V3=V2.
III. Beginning at point 3, a second heat input Q2 is added at constant pressure as the piston moves outward from V3 to V4. Fuel had begun to burn at point 2, and burning is complete at point 4. The total heat input is QIN=Q1−Q2 . The expansion work from 3 to 4 is W34.
IV. After the hot compressible fluid (combustion gas) expands from point 3 to point 4, it continues to expand to V5 without further heat input. The power stroke is complete at point 5. In our engines, point 5 is at a substantially greater volume than point 1. The expansion ratio is defined as RE=V5/V2 and exceeds the compression ratio by the factor A=V5/V1. A is called the Atkinson ratio. It is equivalent to A=RE/RC. The work from 4 to 5 is W45.
V. In this final step of the General Cycle, the engine returns from point 5 to the starting point 1 of
This recharge Step V is inherently irreversible and represents a departure from the fully reversible cycle model as explained in the referenced article by applicant Rogers. In the fully reversible case, this portion of the cycle is assigned two steps: first, a reduction of pressure at constant volume, and then a reduction of volume at constant pressure, to return to the starting point of the closed cycle. Opening the cycle as described here causes a loss of work against the atmosphere. The work against the atmosphere, WATM, is negative. The total work of this cycle is W=W12+W34+W45+WATM. The efficiency of the cycle is obtained by dividing the total work W by total heat input QIN.
We caution that while the above explanation of the General Cycle is of great benefit for understanding our invention, it represents a particular example and only approximates processes that may occur in a real engine built according to the invention. One may, for example, program the rate of heat input Q2 so as to restrain the maximum gas temperature (rather than maintaining constant pressure as described above) and thereby prevent formation of nitrogen oxides by nitrogen and oxygen molecules present in the combustion gas. Such a useful variation from the General Cycle should be understood to fall within the scope of our invention.
The present invention achieves its major benefits by making use of the unique features of the General Cycle. It is noted that the commonly known internal combustion engine cycles, Otto, Atkinson, Diesel, Brayton and the dual cycle are encompassed by the General Cycle as special cases of it. It is common in the art to refer to a cycle as the least general designation which encompasses all of the cycle's activity. For example, the Otto cycle is a special case of the Atkinson cycle, in which the Atkinson ratio is 1. An engine by that design is referred to as an Otto cycle engine, not as an Atkinson cycle engine. We use the terminology of the General Cycle in a similar way. Our General Cycle engines make use of all of the features of the General Cycle, and do not fall into a special case or category in which one of the other cycles may be the more appropriate terminology. Therefore, we can expect an engine using the General Cycle to use, at least in some fashion, two heat inputs, the first at substantially constant volume, and the second in which the fuel is metered at such a rate that the pressure is held substantially constant, up to the point of fuel cut-off. At least that is the idealized view of what the engine is doing; in the reality of a physical engine the piston is always moving and there is never actually an exactly constant volume, and likewise there is not a perfectly constant pressure. Much of any ideal cycle's description and computations are approximations to the real physical system. These ideal constructs aid our understanding and allow us to achieve good engine designs with reasonable effort.
The General Cycle as presented above does not include a constant-temperature step. However, as currently practiced, the General Cycle may now be extended to include a constant-temperature step. We call this the Temperature Limited General Cycle. All aspects and examples of our invention can be operated in a fashion to include a constant-temperature process with only minor adjustments being required in the fuel injection programming. Near the conclusion of the Detailed Description, we will present a full description of the Temperature-Limited General Cycle. Any engine construction that we may describe herein regarding our invention may optionally include a constant-temperature process as later detailed in the presentation of the Temperature-Limited General Cycle. This cycle has substantially the same sequence of thermodynamic steps as before with the addition of a constant-temperature portion during the last part of heat input.
The General Cycle has been described above as a reversible thermodynamic cycle operating on a perfect gas as the working fluid. This cycle has been used as a design basis for practical engines having improved performance and efficiency in many applications. Engines based on the General Cycle have two preeminent features: (1) control of maximum gas pressure, and optionally control of maximum temperature, as desired for best operation at each power level, and (2) compression ratio and expansion ratio are chosen to obtain best performance within design constraints of any particular application, as may be most desirable, in a great many applications. We will show below that the expansion ratio of an internal combustion engine is the first determiner of efficiency, and that in any particular engine design having a selected expansion ratio, there is a corresponding compression ratio needed to achieve the best engine efficiency. We will further show that for any such particular engine design with a selected expansion ratio, there exists a maximum achievable efficiency. This best efficiency is obtained by use of a certain optimum compression ratio.
By application of the principles we have discovered, one can obtain optimal engine designs for a great variety of applications. One may for example choose to design an engine to achieve a desired level of efficiency, such as 60% brake efficiency, or a design may be selected that provides the best efficiency within certain design constraints such as a desired level of specific power. Finally, we will show that for a great many optimal engine designs utilizing the General Cycle model, the best combinations of compression ratio and expansion ratio fall within certain defined boundaries.
Our design principles and conditions for optimal engine design will be illustrated through several specific examples—two will be two-stroke engines of exceptionally high efficiency and a third example will be a four-stroke engine that may be more suitable for use in large trucks or off-road equipment. We have found that a two-stroke, direct-injected engine generally working in accordance with the General Cycle is superior to other engines regarding the combined properties of efficiency, power density, and ease of construction. This fact is illustrated by the first two engine examples we will present. In a third example engine of a four-stroke design, we will show how our invention may also be applied to construct an engine suited to a particular purpose that is highly efficient and that also has good power density, or specific power, in four-stroke operation.
Our invention concerns the application of General Cycle principles and other conditions to the construction of efficient engines. We will show how they may be applied in novel, high-efficiency engine constructions.
Now, in a first instance, we have found in our work that a practical upper limit of efficiency exists for internal combustion engines of our design. For our engines of most efficient and practical design, best efficiency lies in the general range of 50 to 60 percent brake efficiency or may be even higher, depending on the fuel used. In order to obtain an optimum brake efficiency of approximately 60 percent or greater the following inequality must be satisfied:
AR
C≥36.33+8788 e−0.375 R
For these highly efficient engines, the most desirable values of compression ratio, RC, are in the range from 19 to 30. The design property of Inequality 1 determines highly desired values for ARC, which is the product of Atkinson ratio, A, and compression ratio, RC, and which is also equal to the expansion ratio RE. The following Table 1 illustrates minimum values of ARC satisfying the Inequality 1 for whole number compression ratios from 19 to 30.
Referring to
AR
C
=R
E≥36. (2)
A whole number simplification of the narrow band range of preferred expansion ratios is between 36 and 44, inclusive, as shown in
Although deviations in construction of a practical engine which do not quite satisfy the original inequality may result in an engine with slightly less efficiency than 60%, it will be apparent to those skilled in the art that such an engine would still be highly efficient, and would exceed the efficiency of any practical engines known heretofore. Therefore, it should be considered that any such engine making use of the features and theoretical principles in its design and construction as herein set forth falls within the scope of our invention, regardless of the actual efficiency. Moreover, any engine which substantially approaches the design constraints herein set forth also falls within the scope of our invention.
We will now describe example constructions of engines designed in accordance with the principles that have been presented. In doing so we will describe per example only one cylinder and its accompanying structure, but it will be appreciated by one skilled in the art that engines are commonly composed of multiples of such similar cylinders and parts, and such constructions are within the scope of our present invention.
We will now show a preferred engine construction that uses piston motions to input a fluid or gas such as air into the engine, and to compress and expand the fluid or gas as performed in the General Cycle. This particular example is presented in
During the recharge portion at the end of each cycle and before the beginning of the next cycle, piston 24 is in a position outward from intake port 26 so that the intake port is in communication with cylinder volume 22. A fluid supply means for supplying working fluid to cylinder volume 22 is provided. As an example, a compressible working fluid, otherwise known as a compressible gas such as air is introduced into cylinder volume 22 through a fluid inlet means for admitting the fluid into the cylinder volume through, for example, intake port 26. This fluid or gas is obtained from a fluid supply 33. The fluid supply 33 may be at atmospheric pressure, or may serve to pressurize the fluid, as is common for example with a turbocharger. The fluid flows from fluid supply 33 to a reservoir 34, then through intake port 26 into the cylinder volume 22. Reservoir 34 is external of the cylinder and other engine components such as a crankcase. An optional check valve, such as a reed valve 50, may be placed between the fluid supply 33 and reservoir 34. This check valve can optionally serve to prevent fluid from flowing back toward the fluid supply 33 as the piston moves outward, reducing the back volume 25.
As the piston moves outward, fluid in the back volume 25 is forced out through intake port 26 and a back port 37. The back port 37, which provides for final discharge of fluid from the back volume 25, may be either situated in the end portion of the engine body 27 or adjacent to the shaft bearing 36 in bearing housing 38, as shown in
Additional parts connected to the combustion chamber 23 are an exhaust port 40 with a valve 41, and an injection means for adding fuel to the compressed fluid, such as fuel injector 42. Port 40 with valve 41 form a closable opening to selectably permit transfer of the fluid out of the cylinder. The initial pressure, P1, of the engine can be controlled by the inclusion of an exhaust flow pressure regulator 39 that regulates flow from the exhaust port. By maintaining pressure in cylinder volume 22 during scavenging, the operational characteristics of the engine may be selected. Pressure regulator 39 may be variable for selecting preferred operating conditions during engine use.
A heat input means for increasing the internal energy of the fluid in cylinder volume 22 is provided. In this embodiment the heat input means includes a fuel supply means for adding fuel to the fluid. As an example, this may be an injection means for transferring fuel into the cylinder volume, such as fuel injector 42 which receives fuel from a fuel supply system 43. This increases the internal energy of the fluid by combustion of the injected fuel. The heat input means may be controlled to add heat at a controlled rate, particularly both to raise the internal energy of the fluid to the desired peak pressure P3, and to maintain that pressure for a controlled period of time.
The beginning of a cycle as defined here occurs at the time that the valve 41 is closed in the exhaust port 40 and the piston then begins to compress fluid in the cylinder volume 22. However, this does not occur at the time that the piston is near to the far outward position, called bottom dead center (BDC). Rather, the piston 24 moves inward from BDC with the exhaust valve 41 open until a position is reached where the cylinder volume 22 has been reduced by a factor of 1/A from the substantially greater value V5 referred to above in describing the General Cycle and further described below. A is the Atkinson ratio. (In the present example, A has a value of 1.4 and the desired compression ratio is RC=27.)
At the time that the valve 41 is fully closed, the value of cylinder volume 22 is V1. All valves are closed and compression of the fluid, gas or air in the cylinder volume 22 begins as piston 24 continues to move inward toward top dead center (TDC) position, which is the point of least cylinder volume referred to as V2 in the above General Cycle description. This least cylinder volume is substantially the volume of combustion chamber 23. As the piston 24 arrives at substantially the TDC position, heat Q1 is selectably added to the fluid in the cylinder volume 22 (presently equal to the volume of combustion chamber 23) by the injection and burning of fuel as described by General Cycle Step II. This process continues for a short time, initially controlably raising the fluid to a desired maximum pressure P3 and associated temperature T3 at a near-constant-volume condition. For a brief additional time as the piston moves outward, heat Q2 is controllably added as required and in a fashion to maintain substantially the constant pressure P3, as described in Step III of the General Cycle. Then fuel cutoff occurs. At fuel cutoff, the cylinder volume 22 will have increased in volume to a value V4 as described in the above General Cycle description. The heated gas, at a very substantial pressure, drives the piston farther thus sending power via the piston shaft 28, through power linkage means 30, and to the power transfer means 32. This continues until the piston reaches an outward position approaching BDC, at which point exhaust valve 41 opens to discharge burnt gases from the cylinder volume 22. At the effective time of valve 41 opening, the volume of cylinder volume 22 is substantially equal to V5. Shortly after, the pressure in cylinder volume 22 falls below the pressure of the fluid in fluid reservoir 34. As the piston 24 continues to move outward, it uncovers intake port 26. Then a fresh charge of fluid displaces remaining burnt gases in the cylinder volume 22 and fills the cylinder volume 22 with a fresh charge of fluid. The initial pressure in cylinder volume 22, provided by the fresh charge of fluid, is the intake pressure, P1. This pressure is controlled by exhaust pressure regulator 39. The replenishing of the cylinder volume 22 with fresh fluid continues for a length of time while the piston 24 completes its travel to BDC and where it then reverses direction, and covers the intake port 26 again by its inward motion. The valve 41 remains open for a further time as the piston continues to move inward. The valve 41 closes at the point where the cylinder volume 22 has returned to the value V1. This is the point of beginning of a new cycle.
This two-stroke engine operating at an effective intake pressure P1=115 kPa (1.15 bar) and having a compression ratio of RC=27 has excellent fuel utilization for a broad range of renewable and fossil fuels. It gives good specific power (i.e., the power density in hp/liter) and substantially 60% brake efficiency or greater, depending on the fuel that is used. The engine has the following operating characteristics operating on No. 2 diesel fuel (ASTM D975-19a, 2-D (S-15) at 70% of stoichiometric mixture:
As an example, this small engine has dimensions of:
At 1200 RPM, a mean piston speed of 8.0 meters per second.
Other fuels are also being evaluated:
The above small engine shown in
Fuel flexibility is an important benefit of our high-compression, high-efficiency engines. Except for changes in fuel injection means, the engine of
An especially preferred engine for cogeneration of heat and electric power, which operates generally in accordance with the General Cycle, and satisfies Inequality 1, is shown in
An optional ceramic face 18 is applied to each of the pistons 2. This ceramic material is applied by plasma or flame spraying or other method and in a sufficient thickness to substantially reduce heat transfer from the hot gases to the piston bodies. The combination of the ceramic combustion chamber 16 and the ceramic piston faces 18 is an important optional aspect of our invention as it greatly reduces energy loss by heat transfer.
At the time that the valve 6 is fully closed, the total operating volume is the value of cylinder volume V1. In the first portion of the engine cycle, the pistons 2 move toward the partition 14, approaching its surfaces 13 very closely as the pistons reach their top dead center (TDC) positions. In so doing, they compress a compressible fluid or gas 19 into the combustion chamber 16. This fluid 19 may comprise oxygen, air, combustible gas or vapor, or any combination of suitable gases. The TDC position is the point of least cylinder volume referred to as V2 in the above General Cycle description. This least cylinder volume is substantially the volume of combustion chamber 16. Heat is introduced into the combustion chamber 16 by a heat input means configured to increase internal energy in the fluid by injection of fuel through fuel injector 7, which fuel almost immediately commences burning in cooperation with the compressed fluid 19, thus forming a combusted gas at high temperature and pressure. Heat Q1 is selectably added to the fluid 19 in the cylinder volume (presently equal to the volume of combustion chamber 16) by the injection and burning of fuel as described by General Cycle Step II. This process continues for a short time, initially controlably raising the gas temperature and pressure to a desired maximum pressure P3 and associated temperature T3 at a near-constant-volume condition. (The volume of fluid 19 is nearly constant during the heat addition Q1 because this first heat addition occurs when the piston is near to TDC and it is moving very slowly.) For a brief additional time as the piston moves outward, heat Q2 is controllably added as required and in a fashion to maintain substantially the constant pressure P3, as described in Step III of the General Cycle. Then fuel cutoff occurs. At fuel cutoff, the cylinder volume will have increased in volume to a value V4 as described in the above General Cycle description. In the next portion of the cycle, the pistons 2 move outward, transferring energy to crankshafts 4 by means of connecting rods 3. This is the power stroke of the engine. The two rotatable crankshafts are mounted in relation to the engine body, and the connecting rods or linkages between each crankshaft and its associated piston drive the pistons or extract energy from the movement of the pistons. The crankshafts are timed to advance the pistons at substantially the same time. The power stroke ends as the pistons 2 draw near to uncovering a fluid inlet means for admitting fluid into the cylinder volume, in the form of intake ports 9 in the walls of the engine body. The exhaust port with valve 6 opens at the end of the power stroke, and combusted gas is discharged from the cylinder volume. This cylinder volume is composed of combined volumes 12 and combustion chamber 16. The combusted gas is discharged through an exhaust manifold system 8. Shortly afterward, the pistons 2 pass outward sufficiently to uncover intake ports 9. While the pistons 2 are outward past the intake ports 9, fluid 19 enters through the intake ports 9 and displaces remaining burnt gases within the volumes 12 and combustion chamber 16. As the pistons 2 reach BDC, they reverse their direction of motion and begin to move inward again.
A third portion of the cycle comprises the operation of the engine between the time of beginning of inward motion of the pistons 2 and the time at which the engine again begins to compress fluid for a new cycle. During this time interval, the pistons move a substantial distance inward. The end of the interval is defined by the effective closure of the valve 6. A key aspect of our invention concerns the positions of pistons 2 and the total operating volume of the engine at the times of opening and closing of the valve 6. The total operating volume is equal to the volume of the combustion chamber 16 plus the combined volumes of the two volumes 12. This total operating volume will now be referred to simply as “V” with a designating subscript that indicates the value of V at a particular point in the engine's cycle. With reference to the P-V diagram of
V1/V2=RC
V5/V1=A
And ARC≥36.33+8788 e−0.375 Rc as has been discussed in detail above.
This second preferred engine construction is considered to be of great value for use in distributed power generation. The engine is imagined to be coupled to one or more electric generators of any desired type. In addition, the waste energy is to be collected at available locations. Approximately 60% of the energy in the engine fuel will be delivered as work to the electrical generator(s). Of that work, as much as 96% to 98% may be converted to electrical energy. The waste energy comprises approximately 40% of the energy in the engine's fuel. A portion of that energy may be collected in the form of high-quality heat. The efficiency of collection and transfer of this heat may be in the general range of 75% to 80%. The heat is referred to as being of high quality because it can be collected at a very substantial temperature, in the general range of 100 degrees Celsius to as much as 300 degrees to 400 degrees Celsius. High-quality heat has great practical value for heating, producing hot water or steam, and for industrial process heat. Combining the two system efficiencies, the engine's efficiency of producing electrical power with the efficiency of collection and use of heat, provides an overall system efficiency of approximately 90%. In this fashion, our invention can be of immense value in reducing dependence on fossil fuels, or fuels of any kind. Our engines may use any of several suitable fuels such as natural gas or biomethane, dimethyl ether, methanol, diesel fuel, gasoline, or a combination of fuels. Some of these fuels may be obtained from renewable as well as geologic sources. Basic physical properties and design parameters of the above example engine are:
As mentioned, our cogeneration engine described above can provide both heat and electricity with a combined efficiency of 90% or greater, and causes no net increase in atmospheric greenhouse gas (e.g., CO2) in its operation when using a renewable fuel. This engine combined with a synchronous generator has a fuel consumption of approximately 0.148 kg/kWh when operating on ultra-low-sulfur diesel fuel. Electrical generation is at approximately 57% efficiency. Thus, our new engine technology represents a great advance toward reduction of global warming and climate change.
It is to be appreciated that the foregoing descriptions of engine configurations are only representative of the engines of our design. There are many parameters which are open to selection by an engine designer to accomplish the end goals needed for a particular application. Our designs and methods of operation are particularly focused on efficiency, and how to get maximum energy from the fuel expended. That is, to convert the most of the fuel possible into useful work, and besides, also to capture waste heat at a useful temperature.
Many elements of engine design are trade-offs. For instance, power and efficiency are commonly recognized as a trade-off. An engine having very high specific power will often have lower efficiency. Conversely, an engine having higher efficiency will often be thought of as having low specific power. Nevertheless, the efficiency of engines of our design is not solely due to sacrificing specific power. Following the principles laid out for engines of our design results in optimally efficient engines for a selected or desired power.
Our research confirms a longstanding belief in the art that greater efficiency of an engine is had by using a high compression ratio. But our research also shows that there are limits to what is practical. These limits depend on the peak pressures and temperatures involved.
The General Cycle likewise refers necessarily to a construction and operation that has an Atkinson ratio greater than 1. (Recall that the expansion ratio, RE, is related to the compression ratio, RC, and the Atkinson ratio, A, by the expression, RE=A RC. Thus, having an Atkinson ratio greater than one is equivalent to having RE>RC.) Our research also confirms a longstanding belief in the art that engines having Atkinson ratios greater than 1 can produce power at greater efficiency.
But our research also shows that there are practical limits to efficiency gains by employing the Atkinson ratio by itself. These limits depend on the compression ratio, fuel injection timing, friction and other energy losses, and specific power. These are reasons that an engine designer does not just use a high Atkinson ratio and expect a great increase in efficiency. Selection of the correct Atkinson ratio is a major subject of the present invention. There are physical limits which confine the design of engines. These include peak pressure, temperature, and specific power. They place limits to using increasing compression ratio and expansion ratio to gain efficiency.
There are also business reasons contrary to pushing the limits of efficiency. They relate to the costs of production and operation. Many of these considerations have to do with the intended use of the engine. Engines used in transportation are limited to lower weight and fitting into regular cubic form factors. They also need to have high specific power. Stationary power generation engines do not have such tight constraints, but still they must be within the limits of a cost/benefit ratio as given by fuel costs and operational costs as well as expenses for engine manufacture and installation or site costs.
All of these things, physical limits and business reasons, have bearing on whether or not a particular design is “practical.” For the purposes of deciding which engines are practical, we establish reasonable limits of peak pressure, temperature and specific power, and compare them judiciously with increases in efficiency. Noting that efficiency does not always increase with increasing compression ratio and Atkinson ratio, but moreover that these are limited by reasonable upper limits of peak pressure and temperature, and declining specific power, we selected an efficiency goal of 60% to 65% brake efficiency and strove to define designs that would be practical and yet have high efficiency.
We also speak of the “optimum” or “optimal” efficiency engine. This does not necessarily mean the absolute highest efficiency, but rather the best efficiency one may achieve while maintaining certain other design goals such as a minimum requirement for specific power—that is, power per unit of weight or per unit of engine displacement.
There are always tradeoffs in engine design. It is well known in the art that there is a tradeoff between specific power and efficiency. One may at times get a higher engine efficiency by decreasing equivalence (the fraction of stoichiometric fuel added). But this lowers the power output (and the specific power). Efficiency can also be improved by using a higher compression ratio, but this is expected to lead to higher peak pressures and temperatures. Higher pressures require a stronger structure to contain them. Higher temperatures invoke more parasitic heat loss and create undesirable emissions. Another way of increasing efficiency is to design an engine which is over-expanded in the power stroke, that is, it has a high Atkinson ratio. Not only does this lead to a loss of specific power, but it also increases power losses due to the effect of friction from a longer stroke.
We have found that a lower limit of ARC (or RE) (for obtaining 60% efficiency under practical conditions for a General Cycle engine) of just slightly over 36 is applicable to engines having compression ratios of approximately 26 to 30. Note that in the graph of
On the lower end of our chosen spectrum of compression ratios, to meet the efficiency goal of 60%, the expansion ratio becomes impractically high, with a compression ratio of 19 requiring an expansion ratio of approximately 44. A compression ratio of 19 is found to be a practical lower limit for achieving an efficiency of 60%. Of course, most common engines operate at compression ratios well below 19, which is why they have no hope of having efficiencies anywhere near the above selected efficiency goal of 60%.
The great value of our engine design work is that with all of the unique features of the General Cycle, we have discovered that there is a definite paring of values of each compression ratio with a narrow band of corresponding values of ARC (which is equal to the expansion ratio, RE) to obtain the overall optimal engine performance. It is not just that more Atkinson ratio is better, but rather that there is a specific range of values of Atkinson ratio and expansion ratio that result in high efficiency within the practical constraints of other conditions.
We establish a lower limit of values of the relationship of RC to ARC in order to achieve a target efficiency of 60% to 65% brake efficiency. This efficiency goal is reasonable, and has been suggested in the art as the theoretical upper expectation that may be possible for an internal combustion engine to achieve. However, engines in the prior art have not achieved an efficiency of 60% to 65%. It is well known that most actual internal combustion engines operate in the range of 20% to 40% efficiency. Large industrial engines and ship engines have the best efficiencies known to date, and they obtain 45% efficiency to the highest reported value of 55% efficiency. Thus, a goal of 60% efficiency or better is exceptional. The reason that we do not see commercial engines in the 60% efficiency range is that until the present invention they have not used the General Cycle with the unique paring of RC to ARC that we have developed.
Our unique application of the General Cycle is the use of two heat inputs in the manner described in combination with an optimum pairing of compression ratio with Atkinson ratio. This novel combination of design conditions creates a high-efficiency environment that has not been known in the prior art. This operating environment is key to the great benefits obtained by the practice of our invention.
We believe that a particular construction of parts is not necessary to practice our invention—rather, principles, features, and limits of engine design are defined. We have, however, described three example constructions. These constructions have those particular features which are useful in completing the goals of engine design according to our invention. The specifics of engine construction which are unique and which have utility in helping to achieve the operating characteristics of our invention will continue to be delineated. However, please note that this invention is not primarily a mechanical construction. Any assemblage of parts capable of performing the functions of the General Cycle and which includes our unique paring of RC and RE (or ARC) is within the scope of our invention. Further, we note that the specified goal of 60% to 65% brake efficiency is not necessarily a requirement of our invention. The brake efficiency of an engine includes losses occurring through heat transfer and friction. These losses are subject to many physical construction parameters. It may well be that an engine gets better or worse brake efficiency depending on the specifics of construction and how those details affect energy loss such as caused by heat transfer and friction. Nevertheless, we believe that for any practical manner of engine construction, optimal efficiency shall be obtained for that construction by following the basic ratios and operations obtained by following the principles of our invention.
We will now show how the features and principles contained in our invention can be used effectively to construct engines having the best possible efficiency for a great variety of circumstances including engines having efficiencies both below and above 60 percent efficiency.
Starting at an Atkinson ratio of 1.0, as one moves left along the constant RE line 54, one encounters lower compression ratios, RC, higher Atkinson ratios, and higher efficiency. But increasing efficiency only exists up to a point, reaching a peak value at 56. Then, moving further left, lower compression ratios and higher Atkinson ratios do not enjoy the benefits of greater efficiency, but rather, efficiency decreases.
The stroke length of an engine increases in a definite way with increasing expansion ratio, RE. (Remember from the description of the General Cycle that the expansion stroke begins at point 2 and ends at point 5. These points correspond to cylinder volumes V2 and V5, and the expansion ratio is defined as RE=V5/V2. Thus we see that the expansion ratio contains the stroke length in its definition.)
Since for each point in each of the bands, including the ridgeline value of peak efficiency, we know the compression ratio, RC, and the expansion ratio, RE, thus we also know the Atkinson ratio, A. For the first time, we have discovered not only that there is in fact an optimum Atkinson ratio, but also what that optimum Atkinson ratio is for each ridgeline RE-RC pair.
Now the optimum Atkinson ratio does vary, as has been shown, based upon a host of engine design parameters. The peak efficiencies vary, and may occur both above and below the before mentioned efficiency goal of 60% or more. Nevertheless, we have demonstrated how to design an engine having optimum efficiency for its size and other design conditions. Note that for most efficiency bands, the peak efficiency for a given expansion ratio, RE, is found to be at an Atkinson ratio greater than 1.0. The exception is in engine efficiency family 86, at the bottom band, which represents an RE of 25. For this selected engine type, at this expansion ratio, the most efficient engine is at an Atkinson ratio of 1.0. The small-engine-ridgeline 88 impinges on the A=1.0 dashed line between the expansion ratios of 25 and 30.
Before continuing on to explore applications of the information presented in
We have shown that the principles of our invention have wide application for many types and purposes of engine design. Therefore, we will now show how selection of the expansion ratio with its associated best “ridgeline value” of compression ratio may be utilized to obtain an optimal engine construction for practically any desired application, and we will then present an example engine design illustrating this design method. Table II shows a range of optimal (ridgeline) pairs of expansion ratio RE and compression ratio RC, over the ranges of greatest value and interest, with expansion ratios ranging from RE=25 to RE=50, and we will show how varying of the RE-RC pair influences the properties of an engine whose other features are held constant.
Referring now to Table II, the table presents performance characteristics of one particular engine construction (a very efficient opposed piston, two-stroke engine of substantial size, bore B=187 mm, adhering to our design principles) for which the optimal pair of expansion and compression ratios is varied while the cylinder bore, the combustion chamber size and configuration, and many other engine properties, remain constant. We see from the second and fifth columns of the table that increasing the expansion ratio results in higher engine efficiency. This relationship between RE and efficiency is plotted in
In
A large-engine ridgeline 96 is plotted on this graph for an engine having a cylinder bore dimension of 13 inches (0.330 m) and having RC values even lower comparable to those of RC as used in
Of particular note is that the Atkinson ratio increases as one moves left (or counter-clockwise) from the diagonal A=1.0 line 90. Thus, the ridgelines (line 96 and line 97 and all such lines between, not shown) only contain points for Atkinson ratios greater than or equal to one. At their lower ends, ridgeline values impinge on an Atkinson ratio of A=1.0. Our preferred engines are ones that achieve higher efficiencies than these by utilizing Atkinson ratios of between 1.1 and 1.8. Further, they are engine designs that will generally fall within the enclosed space 95 of the graph of
We caution that while
We will now show that an optimum value of compression ratio, falling generally within the preferred bounded area of
Where RC is the compression ratio providing maximum efficiency
The map shown in
This formula produces RC and RE pair values closely corresponding to ridgeline values in the area we have studied. Thus, for an engine of particular engine parameters of size, construction, piston speed, maximum pressure, fuel type, and equivalence, one can use this Equation 3 to predict the ridgeline pairs of RC and RE as conditions are varied.
Consider that the object of the present invention is to teach methods for design of engines of exceptionally high efficiency for many applications. To this point in our description, we have defined the optimum RE-RC pairs for obtaining generally the best possible efficiency for any engine built according to the principles and methods of our invention. Of course there will be many applications for which the best and most practical design for the application may include small deviations from the optimum RE-RC pairs as have been defined. In referring to the details of
wherein ε is a small fraction such as 0.1, and X has the value,
The maximum compression ratio included within the enclosed space 95 of
The ridgeline of the large engine shown in the upper plot in
It may be necessary to modify the General Cycle operation in some situations for the purpose of limiting NOx formation. This may be accomplished by controlling gas temperature as well as flame temperature in the cylinder. Flame temperature can be controlled by fuel selection and optional additives, operating at lower equivalence, and premixing some of the fuel with the air input. An inherent characteristic of our invention in all of its aspects is that the heat input Q2 is controllably added so as regulate chamber pressure. In the General Cycle mode of control, the heat input is regulated to maintain a constant pressure up until the time of fuel cutoff. This is a very practical method of control that can achieve the highest possible efficiency of an engine. On the other hand, the mode of control of heat Q2 input may be selected to be in any way desired by the engine operator, and in some circumstances it may be desired to extend the period of heat input to some degree. This will always result in a loss of efficiency. A specific method for lowering of the bulk gas temperature is by dividing the fuel input into three parts, including a third portion of heat being added at constant temperature, as described below—this will lower efficiency and thus would only be used where a lower peak gas temperature is required.
Referring to
The above reversible processes and their associated perfect gas relationships provide sufficient information to derive an equation for ideal efficiency of an engine having both limited pressure and limited temperature. With suitable substitutions of variables, the following equation for efficiency is obtained. All variables in the equation are dimensionless:
Wherη is the ideal efficiency expressed as a dimensionless fraction,
While the above Temperature-limited General Cycle is presented as a reversible, closed cycle operating on a polyatomic ideal gas, we have found that the equation may be profitably applied to analysis of real engines by recognizing that Steps VI and VII may be considered to be replaced by steps that exchange exhaust products for a fresh charge of air or other suitable working fluid. The calculated efficiency may then be adjusted to account for energy losses in the real engine being modeled.
In our research we have determined that the use of the General Cycle is critical in obtaining peak values of efficiency as we have described. In particular, the General Cycle has two heat inputs. (Or three heat inputs as in the case of the Temperature-limited General Cycle.) The first heat input is at substantially constant volume, raising the pressure to a maximum value. The second heat input follows the first heat input as the volume starts to increase, and continues until fuel cut-off. The limitation on how much heat (or fuel) can be added in total is the stoichiometric fuel value modified by the equivalence. This specifies the total amount of fuel that can be added, and thus the amount of heat produced. In our research we have found that the highest efficiencies are obtained by the second heat input (including all heat inputs after the first heat input) being at least 20 percent or greater of the total heat input. For this 20 percent figure, we consider all heat added after the first heat input to be counted in the second heat input, whether or not a constant pressure is maintained. The value of 20 percent or greater of the total value for the second (including any remaining heat inputs) is universally applicable in our engine designs in obtaining the optimum efficiency.
In some cases, it may be advantageous to design an engine with a somewhat lower efficiency than 60% in order to have a lower engine cost. This approach to engine design will be illustrated by our third engine example, described below
A third physical embodiment of our invention is shown in
It will be appreciated by one skilled in the art that in order for crankshaft 110 to rotate that crankshaft arms 110a and 110b must be in different planes, as viewed from the end view in this diagram. Accordingly, we have chosen our convention to be that crankshaft arm 110a, connecting rod 114, lever 118, piston linkage 122, piston 106, and cylinder bore 102 are in a foreground plane. Crankshaft arm 110b, connecting rod 116, lever 120, piston linkage 124, piston 108 and cylinder bore 104 are relatively in a background plane. This offset distance is sufficient for crankshaft arms 110a and 110b to rotate about a main journal 126 of the crankshaft in a conventional manner, but to have an overlap in the front ends of the cylinder bores, as will be described.
Crankshaft arm 110a supports crankpin 128, upon which connecting rod 114 pivots. Connecting rod 114 is in turn connected by a bearing 130 to lever 118. The lever pivots about a pivot point 134 which is mounted to the frame of the engine so that as the connecting rod moves it actuates the lever to move along an arc as shown at 132. The other end of the lever moves in a corresponding arc 138. This end of the lever is connected by a bearing 136 to piston linkage 122, which is in turn connected to piston 106 by wrist pin 140. Thus, as the crankshaft rotates it drives piston 106 back and forth in cylinder bore 102. On the other side of the engine the parts are similarly connected by joints so that the crankshaft synchronously moves piston 108 back and forth in cylinder bore 104.
As the crankshaft rotates, pistons 106 and 108 are driven to top dead center at substantially the same time, as is shown in
This construction being of a four-stroke design, one can appreciate that there are four parts of the cycle: intake, compression, power, and exhaust. The two pistons 106 and 108 work together in unison to achieve these four strokes, acting on the working fluid in unison. Therefore, the assembly may be considered to be one single cylinder from the point of view of the working fluid. It should be noted that it is common in the art for actual engines to be composed of multiple cylinders. The same is true of this design, and although only one functional cylinder (the complete assembly of parts holding the working fluid) is shown and described, more cylinders may be arranged along the crankshaft, and they would be timed so that the power strokes would provide even power to the crankshaft.
Now it was noted that the present design has cylinder bores that are offset and are partially overlapped in their front end positions. This is unlike common opposed piston engines in which two pistons are aligned in a single straight bore. There are actually two geometric differences in this design compared to common opposed piston engines. Firstly, in this embodiment, the cylinder bores are not aligned in one straight bore, but rather have their front faces offset on the opposite sides of the interposed partition. Geometrically, a projection of the cylinder bores overlap, but the cylinder bores themselves do not touch because they are separated by the partition. Secondly, in this embodiment, the cylinder bores are tilted at an angle with respect to each other. Both of these geometric differences facilitate the connection of the crankshaft to the pistons. This makes possible a relatively compact opposed piston engine with a single crankshaft. The impetus of the inherent compactness of a single crankshaft leads to the present geometry, which may be desirable in certain applications.
Now continuing with a description of the parts of the engine, between the front ends of the two cylinder bores is a partition, denoted generally at 142. The partition is preferably made of steel, except for an optional ceramic chamber portion.
It is apparent that the partition is the functional unit that allows the alignment of the offset cylinder bores to be possible, and yet to provide for open communication through the combustion chamber between the cylinder bores. It further allows for the cylinder bores to be tilted at an angle to each other.
The engine of this embodiment is contemplated to be used as a large vehicle engine, such as a truck engine or an off-road equipment engine. It would presumably have three or more cylinders of the type described. The four-stroke operation is as follows:
The intake stroke begins as pistons 106 and 108 begin to pull away from the top-dead-center position. As soon as there is clearance, intake valves 154 and 156 open to allow an inflow of fresh working fluid into cylinder bores 102 and 104. The intake of fresh working fluid continues throughout the outward movement of the pistons. In keeping with the Atkinson-type function of our engines, the intake valves do not close at bottom-dead-center. Instead, the intake valves remain open for a period of time as the pistons reverse direction and move forwardly from the bottom-dead-center position. Depending on the engine speed, intake pressure and the valve size, this means that the cylinders may continue to fill past bottom-dead-center, or alternatively they may force a portion of the fresh charge of working fluid back into the intake manifold.
At a selected position during the inward movement of the pistons, intake valves 154 and 156 close. This begins the functional compression stroke of the engine. As the pistons reach top-dead-center position, the working fluid is substantially compressed into combustion chamber 152. Because the pistons approach partition 142 very closely at their top-dead-center positions, substantially all of the compressed fluid is then contained in the combustion chamber. In this top dead center position of the pistons, the pistons effectively block communication between the combustion chamber and the valves, which are not located within the combustion chamber.
At this point injector 184 begins injecting fuel into the compressed working fluid. The injection of fuel is controlled to first increase the pressure of the working fluid to a maximum pressure, and then, as the pistons begin to move outwardly, the injection of fuel is controlled to maintain a constant pressure and/or a desired temperature, for a period of time. Then fuel cutoff occurs.
The power stroke of the engine continues as the pressurized working fluid drives the pistons 106 and 108 outwardly, imparting power to the crankshaft 110. Since the location of the pistons at the beginning of compression is inward from bottom-dead-center, and the power stroke continues through the full stroke, this creates the condition for an Atkinson ratio of a selected amount to be employed in this engine.
As bottom-dead center is again reached, exhaust valves 158 and 160 open, and then as the pistons 106 and 108 move inwardly, they drive the exhaust out of cylinder bores 102 and 104. The exhaust valves stay open until the pistons approach partition 142, and then the valves must close to allow the pistons to complete their inward movement. This completes the four-stroke cycle, and the engine is now in the position to begin the intake again.
It will be observed that the combustion chamber 152 is not emptied of exhaust, because the pistons do not sweep the combustion chamber in partition 142. This is actually not functionally different than common four-stroke engines, which have an area of the cylinder into which the piston does not enter, and thus exhaust gasses within the volume of the combustion chamber remain in the cylinder at the end of the exhaust stroke. Thus, an equal or greater amount of residual exhaust gas is left in prior art four-stroke engines also.
This example of the third embodiment of our engine shows that engines of four-stroke design are also suitable for construction after the manner of our engines, and that they may benefit from increased efficiency by employing the same principles as we have previously explained.
A further important construction benefit that is made apparent from this example is that the partition between the front ends of the cylinder bores, which houses the combustion chamber, serves the purpose of connecting the two halves of the opposed piston cylinder, even if the opposing cylinder bores are offset, or unaligned. This is important considering that the offset of the opposed cylinder bore pair is a requirement for their making connection to a single crankshaft. Further, the partition makes possible that the opposing cylinder bores can be angled, or tilted, with respect to each other. Angling the cylinder bores would be far more difficult in prior art opposed piston engines that lack a partition since it would create a space into which the pistons could not travel, and that this space would be a larger volume than any reasonable combustion chamber. Thus, the utility of a partition in opposed piston engines of our design is further emphasized.
As in our first two engine examples, this third engine is particularly well suited for use of a renewable fuel. As this engine is intended for mobile applications, a liquid renewable fuel would be preferred. Suitable fuels include dimethyl ether, methanol, renewable diesel, or a combination of fuels. Some basic physical properties and performance characteristics are presented below for this engine when operating on pressurized liquid dimethyl ether (DME). The engine as described below has three “opposed-piston” cylinders, each cylinder having two offset and angled opposed cylinder bores, two pistons and a wedge-shaped partition containing four valves and a combustion chamber extending through it.
We have now described our invention, including our three representative examples of construction. We again emphasize that the invention is primarily centered in the concepts and principles of efficient engine design, and not in mechanical constructions. We thus set forth the following claims.
This is a continuation-in-part of U.S. patent application Ser. No. 17/160,356 filed Jan. 27, 2021 and claiming the benefit of U.S. Provisional Patent Application No. 62/969,090 filed Feb. 2, 2020.
