DE102017003146B3 - Jerk-crank mechanism, as well as equipped combustion engine. - Google Patents

Jerk-crank mechanism, as well as equipped combustion engine.

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DE102017003146B3
DE102017003146B3 DE102017003146.0A DE102017003146A DE102017003146B3 DE 102017003146 B3 DE102017003146 B3 DE 102017003146B3 DE 102017003146 A DE102017003146 A DE 102017003146A DE 102017003146 B3 DE102017003146 B3 DE 102017003146B3
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crank
crankshaft
engine
jerk
axis
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Auf Nichtnennung Antrag
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Viktor Hammermeister
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    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02BINTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00Other engines
    • F02B75/32Engines characterised by connections between pistons and main shafts and not specific to preceding main groups
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01BMACHINES OR ENGINES, IN GENERAL OR OF POSITIVE-DISPLACEMENT TYPE, e.g. STEAM ENGINES
    • F01B1/00Reciprocating-piston machines or engines characterised by number or relative disposition of cylinders or by being built-up from separate cylinder-crankcase elements
    • F01B1/10Reciprocating-piston machines or engines characterised by number or relative disposition of cylinders or by being built-up from separate cylinder-crankcase elements with more than one main shaft, e.g. coupled to common output shaft
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01BMACHINES OR ENGINES, IN GENERAL OR OF POSITIVE-DISPLACEMENT TYPE, e.g. STEAM ENGINES
    • F01B9/00Reciprocating-piston machines or engines characterised by connections between pistons and main shafts and not specific to preceding groups
    • F01B9/02Reciprocating-piston machines or engines characterised by connections between pistons and main shafts and not specific to preceding groups with crankshaft
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02BINTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B41/00Engines characterised by special means for improving conversion of heat or pressure energy into mechanical power
    • F02B41/02Engines with prolonged expansion
    • F02B41/04Engines with prolonged expansion in main cylinders
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02BINTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00Other engines
    • F02B75/06Engines with means for equalising torque
    • F02B75/065Engines with means for equalising torque with double connecting rods or crankshafts
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04BPOSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B9/00Piston machines or pumps characterised by the driving or driven means to or from their working members
    • F04B9/02Piston machines or pumps characterised by the driving or driven means to or from their working members the means being mechanical
    • Y02T10/14

Abstract

Today's reciprocating internal combustion engine achieves scarcely half of its theoretically possible efficiency by virtue of manifold losses arising due to the inefficient movement of the piston. This invention performs a jerk crank drive and thus allows jerk-crank drive engine capable of Performing the expansion and compression in its cycle up to four times faster, thereby reducing its heat losses by 3/4. The duration of its gas exchange is extended up to 1.5 times, the intake is intensified, the cylinder charge is increased and the discharge is relieved. In the jerk crank drive, the piston (2) by means of bolts (4), connecting rod (3) and coupling (7) is coupled to the engine crank (5) while the connecting rod coupling joint (3-7) from the return crankshaft (6 ) is guided in a circle against the direction of rotation of the engine crankshaft. Thus, e.g. The efficiency of an Otto-Jerk crank-drive engine with ε = 10: 1 compared to a conventional analog engine by 60-87%, and its peak power increases by 2.26-2.63 times. A diesel jerk crankcase engine (ε = 22: 1) gets a relative increase in efficiency of 47-64%, its peak power increases by 2.07-2.31 times. The CO emissions of a jerk-crank engine decrease generally proportional to consumption, the NO emissions are reduced by 10 to 15 times.

Description

  • introduction
  • This invention belongs to the field of periodic thermal fluid energy engines with compressible fluids and is concerned with the efficiency of reciprocating internal combustion engines. It presents a coupling gear, which is to be used as a crank mechanism in a conventional internal combustion engine. Further, this invention relates to an internal combustion engine equipped with such a crank mechanism, which is referred to as a jerk crank engine.
  • It is known that the equivalent energy of diesel fuel requires almost ten times the volume and almost twenty times the weight of the best modern accumulators. This energy density can not reach a modern accumulator. For this reason alone, today the reciprocating engine remains the cheapest engine for driving self-sufficient means of transport. For the combustion engine, there is no equivalent alternative today, so mankind is forced to use it for a certain amount of time. Nevertheless, its low efficiency and its manifold harmful emissions are alarming. Most of its emissions can be largely neutralized thanks to catalyst technology. However, CO2 emissions can at best be made carbon-neutral if the combustion engine uses synthetic fuels derived from hydrocarbons, the production of which requires the entire amount of energy needed from renewable energy sources. It becomes completely carbon dioxide-free, using hydrogen produced from renewable energy sources. However, this technology is not yet available because of low efficiency of the entire chain from the recovery to the conversion of the driving power of the fuel cell according to the so-called WTW analysis method ("Well-to-Wheel"), heavy traction batteries, difficult processing and storage of hydrogen mature.
  • However, carbon dioxide can be significantly reduced by increasing the efficiency of the internal combustion engine. And it has enormous potential for improvement: only a few large combustion engines achieve 70% of their theoretically possible efficiency, but the majority less than 50%.
  • The terms and abbreviations, units of measure and parameters used in the further description are explained in attached Tables 1 to 3.
  • Representation of the problem
  • As you know, the upper efficiency limit of an ideal heat engine is given by the Carnot process. η Carnot = 1 - T min / T Max
    Figure DE102017003146B3_0001
    where is:
    • T max is the starting temperature of the working gases in the process in Kelvin;
    • T min is the final temperature of the working gases in the process in Kelvin.
  • In a real heat engine, the value of Carnot's efficiency can never be achieved because of the inevitable heat losses. However, one can considerably reduce the difference between the theoretical and the actual value of its efficiency, if the processes of the thermodynamic processes are made more effective.
  • The low efficiency of a real heat engine in transforming the heat into mechanical energy (≈50%!) Is due to a very unsatisfactory and untimely conversion process. The basic mechanism of today's engine is relatively simple, sturdy, durable and affordable. However, its construction has not changed significantly since its creation and its potential in today's design is virtually exhausted. His further refinement with usual means is not promising. This is shown by the developments of recent decades, during which the efficiency of the reciprocating engine has barely risen, although it is still far below its theoretically possible level of efficiency. Although car fuel consumption has fallen, this is almost entirely due to the reduction in weight of its bodywork, which is made of lighter aluminum and (or) better aerodynamically shaped. In principle, changes must be made in order to extract the remaining potential, which is approximately 50%, from the reciprocating engine.
  • The requirements for the economy of the engine and its emissions are probably just stricter because of the increased environmental impact. This calls for rethinking and economic development is ripe for this: productivity in mechanical engineering has increased and manufacturing costs can be reduced even as the complexity and quality of the engine increase. Although the present reciprocating engine has been modernized in many ways today, its core - the crank mechanism - still remains an archaic design that prevents its further development. Nowadays it pays to invest in the crank drive to make the engine more economical and cleaner because it is still needed.
  • Beginning in this aspect, however, one must point out the following obvious defects of the ordinary crank mechanism:
  1. 1. Even when operating the engine in the state of its highest efficiency, the expansion of the working gases in the power stroke is much too slow. The main drawback here is that the expansion in the area of the TDC is retarded in its high-temperature phase, where the ratio of the area to the volume of expansion space for a Quadrathuber with a typical compression ratio of 10: 1 in gasoline and 22: 1 in the diesel engine from three to eight times higher than UT. The unfavorable movement of the piston in this phase caused by prolonged contact of the highly heated working gases with the large cooled cylinder walls particularly large heat losses;
  2. 2. the suction without charging is only in the relatively low speed range, which is far from the point of maximum power of the engine, sufficient. As the speed increases further, the duration of the gas exchange decreases, the cylinder filling becomes more and more inefficient and the torque decreases. The maximum power of the engine is achieved only by high rotational speed with considerable manifold losses and thus with enormously low efficiency;
  3. 3. the compression is delayed in its final phase, when the temperature of the fresh gas is highest by the heat of compression, whereby a disproportionate part of the work invested in the compression work is lost as heat to the cylinder walls;
  4. 4. the ejection is very difficult due to the greatly shortened duration of this cycle in the middle and high engine speed range through high throttle losses in limited valve cross-sectional areas, therefore, the motor must spend part of its kinetic energy gained in the power stroke to expel the exhaust gases;
  5. 5. The power stroke to gain more time for discharging and thereby facilitate it is purposely terminated much earlier than the piston reaches its UT, therefore the existing geometric expansion of the engine is not exhausted, which reduces the efficiency of the engine.
  • basic idea
  • As is known, the transferred heat Q between a body and its surroundings is proportional to the heat transfer value α, the area S of the body, the difference in temperature (T1-T2) between the body T1 and the environment T2, and at the time t of heat transfer: Q = α × S × ( T1-T2 ) × t
    Figure DE102017003146B3_0002
    Equation (2) is used in the further illustration of the invention for determining the magnitude of the quantitative divergences of the heat losses of the engines. It is evident that the heat losses of a motor in the working and compression stroke can only be reduced by reducing at least one of its four multiplicands. The heat transfer value α and the temperature difference (T1-T2) are not affected because in this invention it was not considered to replace the materials of the engine or its fuel, therefore, only the area S and the time t come into question. The area S is also not changed in the invention construction, although by their reduction in engine already long positive results are shown. Thus, in the construction of the known opposed piston engines, where the area of the combustion chamber walls is almost halved by missing cylinder heads, significantly lower heat losses have been detected. Unfortunately, its design has not been established because it is not sufficiently mature today. Thus, the only option left is to shorten the duration of the contact of the gases with the walls of the expansion space in the cycle of the engine, which is also easily achieved by increasing its rotational speed. As a result, the duration of the cycle as a whole decreases and the duration of each of its individual cycles is complied with. In reality, such blanket clock reduction in a HM but counterproductive. Although with faster rotation of the motor shaft, the heat losses in the Actual work and compression strokes can indeed be significantly reduced, the total cycle losses at a certain rotational speed, which is within the lower speed range of the engine and is specific to each engine design, greater the savings in these cycles. This is due to the simultaneously disproportionately fast growing loss of ejection, the poorer cylinder filling and the increasing mechanical losses instead. To circumvent this problem with a HM, one would have to speed up the duration of the work and compaction cycle in its cycle, and at the same time extend the duration of the suction and ejection. But these possibilities are not offered by the rigid conventional crank mechanism, which is why the actual efficiency of a conventional engine is so low compared to its theoretical one.
  • The described problems can only be eliminated with a crank mechanism, which gives the piston a type of movement, which can be performed with much higher efficiency every single cycle of the cycle. This basic idea is e.g. in the so-called short-haul (KPM) engines (see [2] and [3]).
  • In these designs, a complete piston stroke from TDC to TDC, as well as in engines with conventional crank drives, is performed at a crank angle of exactly 180 °, but the same time is required for the same rotational speed of the motor shaft. In contrast to conventional engines, however, the piston of short-haul piston engines already recovers over 90% of its stroke in the first quarter turn of its crankshaft from the TDC at almost twice the speed. Such a piston travel is usually timely for the working stroke of a HM and must not be exceeded for the expansion of the working gases in the power stroke when a ZYKLUS economical is sought (the exhaust valve is opened at 135 ° CA with 89.6% of the stroke to facilitate the ejection). However, in the second quarter turn of the crankshaft, the piston of a KPM travels its slight residual travel to UT at greatly reduced speed. Such a type of movement of the piston leads to a professional adaptation of the valve timing of a significant reduction in heat losses in the working and compression stroke of the engine by shortening their duration. The suction is intensified and lengthened at the same time, which requires a better cylinder charge. The ejection is also extended and thereby facilitated. The entire cycle of such an engine is much more economical. However, these short connecting rod crank mechanisms also have the following weak points:
    • 1.The mechanism is difficult to design as a compact uncomplicated plunger construction because of its considerable connecting rod tendency, because the piston threatens to tilt;
    • 2. in a crosshead design of the short connecting rod crank mechanism, which is better suited than the plunger construction, the maximum Pleuelneigung which determines the economy of the engine and theoretically limited with the vertical position of the connecting rod to the cylinder axis, in reality, but not greater than be about 70 °. With a further desirable increase in this efficiency-determining slope, enormous lateral forces are directed to the crosshead and the surfaces of its slideways. In addition, the upward displacement of the crosshead to the TDC by the connecting rod when compressing and expelling at the TDC becomes problematic;
    • 3.The most significant for the economic efficiency of a motor maximum expansion speed in the cycle of a Kurzpleuelmotors can be compared to a conventional engine theoretically only up to about 60% (limit) increase, but this is practically unattainable.
  • Determining the heat losses of a real engine is not difficult in reality because you can measure all its parameters. On the other hand, it is not possible to find out the heat losses of a virtual engine design in this way. Therefore, here only a purely computational comparison of the parameters of the engine design of the invention according to all relevant criteria with the parameters of a base of the same real contemporary engine design, whose data are widely known, made. The differences found in this comparison are intended to illustrate the quantitative improvement in efficiency and other advantages of the invention.
  • For comparison, two completely identical modern internal combustion engines are used here conceptually. In this case, in one of them the currently known centric sliding-crank drive is replaced by the invented jerk-crank drive with the same stroke and its valve timing is adapted to it. Thus, a prototype of a jerk-crank drive motor is created whose base is not the least different from the base of this HM. By such a conversion, the engines to be compared maintain the same compression ratio and the same fuel, the materials of the walls of the expansion spaces remain indistinct and the shape and dimensions of their cylinder spaces do not change, therefore, the parameters α and S in the equation (2) remain unchanged and can be omitted justified.
  • The term "k1" below always denotes the relative size of an element or distance to the crankshaft length of the engine crankshaft (6) (see Table 3 below).
  • The following conditions for comparing the motors are defined and assumed:
    • ✓ that the prototype of the jerk-crank-drive engine is equipped with a crank-drive, which has a distance of 1.25 kl between the axes of rotation of its crankshafts, which is equal to its coupling length and the two crankshafts have equal radii of crankshaft cranking of 1 kl;
    • ✓ that the comparison of all the parameters of the motors is carried out at a constant rotational speed of the motor shaft, which corresponds to the speed and power of the HM, at which it obtains its maximum efficiency;
    • ✓ that the comparison takes place at the prescribed operating temperature of the HM, the temperature of the fresh mixture and the atmospheric pressure are identical and the mass ratio of the mixture is kept at a stoichiometric combustion air ratio (λ = 1);
    • ✓ that the power stroke of the engines to be compared for a piston stroke which corresponds to the 95% piston stroke from the TDC (position of the engine crankshaft in the HM: 148.7 ° CA, in the jerk crank engine: 43.2 ° CA), by the Opening the exhaust valve is terminated and the expansion loss of 4.5% (ε = 10: 1) to 4.77% (ε = 22: 1) is accepted with a corresponding reduction in the theoretical efficiency in the power stroke;
    • ✓ that the temperature of the walls (T2) and the fluctuations in the wall temperature of the expansion or compression spaces of the motors remain congruent during all the processes of the cycle;
    • ✓ that because of the unknown values of the instantaneous heat losses of the engine of the invention, the determination of the instantaneous gas temperatures (T1) in the cylinder of the two engines is carried out as if there were no heat losses during compression and expansion, and only from the base temperature at the beginning of the Process and the instantaneous expansion or compression degree of the gas in the cylinder depends.
  • According to the last assumption, the instantaneous gas temperature in the cylinder of the engines is therefore considered as a pure function of the piston stroke during comparison. As a result, in the equation (2), to simplify the comparison, the expression (T1-T2) also need not be considered. This is justifiable because with completely identical expansion spaces of the engines such an execution determination in comparison alone is to the detriment of the invention. As will be seen later on, in the jerk crank motor, both its full work and compression strokes as well as its most wasteful phases are traversed faster than in a HM. Therefore, in reality, the jerk crank motor has significantly lower heat losses than an HM (in more detail later).
  • Really informative and important in comparing the engines here are firstly the duration of the working cycle and secondly the duration of the loss phase of the compression (the period from the point at which the temperature of the fresh gas through compression exceeds the temperature of the cylinder walls until the piston arrives at the TDC ), because exactly these periods of the invented construction are significantly reduced. These periods are measured here in relative time units (° of the revolution of the motor shaft), which are valid for comparison, regardless of the real rotational speeds of the motor shafts. In reality, however, the absolute heat losses in the cycle of a motor according to equation (2) are proportional to the REAL TIME. And because the cycle time depends directly on the rotational speed of the motor shaft, the absolute heat losses of each motor also depend definitively on its rotational speed. Why then the relative time? The relative time, independent of the speed of rotation of the motor shaft, measured in ° of the revolution of the motor shafts, is only suitable for COMPARING the RELATIVE heat losses of different engine designs. It is needed to be able to objectively compare the processes of a virtual motor with those of a real engine in the proper mode, because one can not measure the duration of the processes in the virtual motor in real time. However, to compare the heat losses in the process of a virtual engine with those of a HM only the relative time is available, which is just as meaningful as the real time in evaluating the relative heat losses of a real engine, because the HM operated in the speed range of its best efficiency is used for nonjudgmental comparison here this relative time is based.
  • the solution of the problem
  • The aim of this invention is to reduce the heat losses of reciprocating internal combustion engines by optimizing the cycle, thereby increasing their efficiency and reducing their harmful emissions. This invention is an alternative to the above-mentioned mechanism of the short-jointed engine, which perfects its advanced characteristics and does not have its mentioned deficiencies.
  • As a basis for a more appropriate crank mechanism of a reciprocating engine is shown in 1 - 1 Mechanism well suited. It consists of two identical crankshafts whose crankshaft radii R and R1 are the same: Crankshaft (5) is embedded with its shaft journals in the bearings (14) and crankshaft (6) with their shaft journals in the bearings (15) of the housing (13). They are rotatably coupled at their crank pin with the coupling (7) whose length K is equal to the distance D. This forms a resounding coupling gear with four-membered kinematic chain. Positioning the coupling (7) exactly on any one of the two Either points (EOP-1 or EOP-2) of the crank shaft, as in 1 - 1, a and 1 - 1, b is shown, then arise only from these two positions out the following alternatives: the cranks to rotate this mechanism either in the same or in the opposite direction to each other. When turning the cranks in the same direction, the coupling (7) remains constantly parallel to the crankshaft axis, the cranks rotate congruent regardless of the distance - like the wheels of a steam locomotive, which are coupled to a push rod (s. 1 - 1, c ). But turn the cranks of this linkage when starting from any either-or-point opposite to each other (s. 1 - 1, d ), now tilts the coupling (7) to the crank pin axis and the uniform rotation of one of the cranks (here, the crankshaft (5) as a motor shaft that turns evenly), gets the rotation of the other (crankshaft (6), the is used as a cylinder crankshaft) a cyclic rotational nonuniformity.
  • In contrast to the same direction rotation of these crankshafts to each other, in which the distance does not affect the function of the rotational movement in the coupling gear, it plays a key role in contrasting rotation of the crankshaft: it determines the degree of rotational nonuniformity of the cylinder crankshaft. The smaller the distance, the more tilted in this mechanism, the coupling (7) to the axis of the linkage and unevenly, the cylinder crankshaft rotates.
  • In the 12 shown constructions two thrusters are based on such a linkage shown. The piston (2) or crosshead (8) are coupled by means of bolt (4) and connecting rod (3) to the crank pin of the crankshaft (6) and are guided in the cylinder (1) or Gleitlagerwandungen (12) in a straight line on its axis. The position of the links of the two crank mechanisms is shown at the end of the power stroke on the piston stroke with a 95% stroke from the TDC, which was generally set for the duration of the working stroke of all comparative engines. As can be seen, the motor shaft (5) needs a smaller crank angle for the same piston travel when the coupling (7) is shorter and more prone: in the crosshead construction 12, a , where the coupling length is 1.25 kl, it needs 43 °, where in the 12, b , which shows a plunger construction, where the coupling length is 1.5 kl, it must cover a crank angle of 71 °. The piston (2) is now greatly accelerated while smoothly rotating the engine crankshaft close to its TDC, but decelerating sharply close to its UT (BDC) much earlier. This positive influence on the piston movement of the coupling tendency is much stronger here than that of the connecting rod in the traditional crank mechanism and even stronger than that of the connecting rod in the Kurzpleuelkurbeltrieb. The acceleration of the piston by the inclination of the coupling and the connecting rod or its retardation by the repeated erection of these links to the crank drive or cylinder axis is added together here and its movement curve is thereby greatly changed.
  • The family of curves of the piston movements of the crank thrusters with different parameters compared to some motion curves of different types of crank gears in 3 clearly shows the advantages of the jerk crank mechanism. Below are the required time periods for the execution of the working cycle with specified 95 percent piston stroke from TDC to degrees of engine shaft rotation (relative time!), Depending on the design and parameters of the crank mechanism with constant connecting rod ratio λ = ⅓ (except short-jointed engine). 1. Sinus curve as a basis for comparison. Theoretical duration of the working cycle 154.2 °; Second Curve of a HM, duration of the working cycle 148.7 °; Third Curve of an HM with displacement of the cylinder axis by 1.0 kl, duration of the working cycle 135.6 °; 4th Curve of a KPM with connecting rod 70 ° (λ = 0.9397), duration of the working cycle 108.6 °; 5th Curve of a KPM with connecting rod 90 ° (λ = 1,0), theoretical duration of the working stroke (duration of the complete downward stroke) 80.3 ° (90.0 °); 6th Curve of a jerk crank drive with a distance of 1.25 kl, duration of the working stroke 43.3 °; 7th Curve of a jerk crank drive with a distance of 1.25 kl, δ = 15 °, cylinder axis is tilted by 15 ° to the crankshaft axis, duration of the working stroke (duration of the complete downstroke) 29.2 ° (80.3 °); 8th. Curve of a jerk crank drive with a distance of 1.25 kl, e = 1.0 and a displacement of the cylinder axis by 1.0 kl from the plane in which the shaft journals of the two crankshafts are located, duration of the working cycle (duration of the complete downward stroke) 24.3 ° (48 °).
  • For one of several feasible variants of the crank drive engine serves as a basis, the known construction of the crosshead motor (s. 2 - 1, a ). By redesigning such a crosshead engine design accelerated expansion and prolonged retention of the piston at the UT is to be achieved. For this purpose (s. 2 - 1, b ), the crosshead (8) by means of the connecting rod (3) and the coupling (7) is rotatably coupled to the crank pin (5/2) of the engine crankshaft (5). In addition, the connecting rod coupling joint (3-7) is rotatably coupled to the crank pin (6/2) of the auxiliary crankshaft (6) which is additionally installed in the motor housing (13). The connecting rod coupling joint (3-7) rotates in the opposite direction of rotation to the engine crankshaft (5) to the shaft journal (6/1) of the crankshaft (6), which keeps it with its crank pin (6/2) on a circular path. During the smooth circulation of the engine crankshaft (5), the crankshaft (6) receives a cyclical rotational irregularity, which causes the desired jerky movements of the piston (2). It is referred to as a crankshaft (6), because it gives the piston the relevant jerky movement in each cycle in the opposite direction to the uniformly rotating engine crankshaft.
  • In this embodiment of the jerk crank motor, the old, proven construction of the crosshead motor is retained. The height of such a 4-stroke engine, measured from the TDC of the piston to the axis of the motor shaft, at a distance of the crankshafts of 1.25 kl is thereby greater by 15 to 20% than the height of a corresponding crosshead motor when the axis of the motor shaft in a Line with the cylinder axis is. This variant is e.g. for stationary engines, large marine engines and similar promising, where space and weight have no directional significance.
  • A more compact version of a jerk crank motor can be produced using a plunger motor as the basis (see FIG. 2 - 2, a ). In the kinematic chain of such a crank mechanism, the piston (2) additionally assumes the function of the crosshead and directs the normal to the cylinder axis transverse forces (F n ), in response to the joint (2-3) of the piston force (F p ) and to Their angled connecting rod force (F r ) arise, through the socket of the cylinder (1) to the motor housing (13). In 2 - 2 B is in the cylinder (1) displaceable piston (2) by means of bolt (4), connecting rod (3) and the coupling (7) with the crank pin (5/2) of the engine crankshaft (5), with their shaft journals (5/1 ) in bearings (14) in the housing (13), rotatably coupled. The connecting rod coupling joint (3-7) is of the crank pin (6/2) of the return crankshaft (6), which with its shaft journals in bearings (15) in the housing (13), contrary to the direction of rotation of the engine crankshaft (5) guided in circles. This mechanism also ensures that the piston (2) moves very unevenly in each cycle and by appropriate adjustment of the valve timing, the periods of different cycles can be made more appropriate in the cycle. The height of such a crank-type four-stroke engine with 1.25 kl distance of crankshafts compared with the height of a corresponding conventional plunger motor, measured from the TDC of the piston to the axis of the engine shaft, will thereby increase by approximately 25%.
  • But if the motor shaft is not installed in a line with the cylinder axis, but installed laterally from the cylinder axis and coupled by a correspondingly offset Zusatzkröpfung to the cylinder crankshaft, increases the height of the two constructions, described in 2 - 1 and 2 - 2 , Not. As a result, the engine is only wider and takes place an extension of the motor along the axis of its motor shaft. However, the return crank (6) must be strengthened, as it is exposed in this case also compressive forces and torques. This is explained in detail in section 13, see drawings in 13 and 8th , Some crank mechanisms that have similar constructions with coupled crankshafts are: DE 8624014 U1 "Piston engine", [7]; DE 60316372 T2 "Internal combustion engine", [8]; FR 900953 "Reciprocating engine with crank mechanism, which interacts with a lever mechanism", [9]; DE 199 42 904 A1 "Internal combustion engine", [10].
  • Realization of the counterrotation of the crankshafts from the either-or-points
  • As already mentioned in section 4, the relative direction of rotation of the crankshaft (6) to the engine crankshaft (5) when starting the jerk crank engine from an either-or point, if the forced run is missing, depending on whether the coupling in the EOP goes through or back, set as desired. The drawing ( 10 ) represents the jerk crank mechanism in both positions, in which the random change of the relative direction of rotation is possible. In order not to leave the starting direction of rotation of the EOP to chance and prevent a jerk-crank drive motor comes to an inappropriate rotational direction state, a device is provided for each of the EOPs, which causes the crankshaft of a jerk-crank drive at start from a EOP constantly run contrary to each other and turn in opposite directions. As opposed to the designation of a synchronization device, which produces a synchronization, this, which causes the mating, divergent genant. Divergator (10) is assigned to EOP2 and divergator (11) to EOP1. The crank handle (6) holds with its crankpin the connecting rod coupling joint (3-7) on the given circle or arc, for which it is exclusively determined in this construction. It is loaded neither with torques nor with compressive forces. It is continuously exposed only to fluctuating tractive forces, which are constantly zero at positions of the crank in the EOPs. Therefore, the design of such a return crankshaft (6) has a lightweight construction and its mass is small in relation to the engine crankshaft (5). To set them in motion when starting the engine from an EOP in the appropriate direction, only their inertia must be overcome. To accelerate them from their idle state to the idle speed of the engine is therefore not problematic at engine start with moderate acceleration and limited mass.
  • The jerk crank mechanism belongs to the group of non-uniformly transmitting transmissions and therefore has a periodically fluctuating transmission ratio between the coupled crankshafts. The degree of fluctuation of this gear ratio depends on the distance of the crank mechanism. If the parameters of the members of the four-bar chain of a jerk-crank drive motor are selected such that the back-crankshaft can rotate in the opposite direction to the engine crankshaft, the gear ratio fluctuating between the engine and the back-crankshaft reaches its extreme values at the EOPs because the EOPs are also the extreme points the translation function of a coupling gear (also in the mathematical sense) are. Therefore, a divergator must have a similar fluctuating gear ratio. For example, an elliptical gear transmission is suitable for this purpose - also belonging to the group of non-uniformly transmitting transmissions - whose extreme value of the transmission ratio coincides with that of the crank mechanism. If one chooses the ellipses the same size, then the center distance is constant and the wheels turn around their ellipse focal points. Also, a gear transmission with conventional cylindrical gears with constant gear ratio involute and a gear ratio corresponding to the extreme value of the gear ratio of the crank mechanism is used as a divergent when the tooth gap is modified accordingly, because in the narrow region of the tooth contact, the gear ratio does not change much.
    10, a shows the divergator (10) in position of the waves in the EOP2 during tooth contact. His gear consists only of a pair of teeth (1), built on the outer side of the cheek of the return crankshaft (6), s. View according to arrow D, which forms a tooth gap. The pinion has only one on the inner side of the cheek of the engine crankshaft (5) set tooth (2), s. Presentation AA. Once per revolution, when passing through the EOP2, engages the tooth (2) of the engine crankshaft (5) in the tooth gap of the teeth pair (1) of the return crankshaft (6) (view D) and prevents the crank strikes back at this point. 10, b shows the divergator (11) in the position of the waves when tooth contact in EOP1. Similarly, it consists of a pair of teeth (1/1), built into the cheek of the engine crankshaft (5) on the outside, which forms the tooth gap (see CC), and only one in the cheek of the crankshaft (6) internally mounted tooth (2 /. 1) (see BB). The tooth contact also takes place here only in the narrow angular range around the EOP1.
  • It is sufficient that the tooth contact takes place only in a narrow angular range of crankshafts around an EOP. The modulus of the teeth is chosen as needed and the translation also works with abnormal limit number of teeth, since the pinion usually has only a single tooth, which only comes in a very narrow circle cutout in contact with the only gap-forming teeth pair of the wheel.
  • As in 10 illustrated, the tooth of the pinion requires a tooth gap, if there is a need to trigger the crank from both directions opposite to the direction of rotation of the motor shaft, such as in a marine diesel, when its shaft is rigidly connected to the propeller shaft, for the return trip with reversal of the motor. If the engine crankshaft already rotates, this mechanism also runs in the area of the tooth contact during idling, because now the conrod coupling joint with the crank handle breaks through an EOP solely due to the inertia forces of the connecting rod and the cranking links. There are also other mechanisms for the predetermined direction of rotation of the cranks of the jerk crank mechanism to each other conceivable. Self-sufficient, electrically, pneumatically or hydraulically operated starting devices for larger engines are also to be considered. It is possible, for example, to use a switch-off device which ensures tooth engagement only when starting and interrupts the tooth engagement after reaching a certain rotational speed, for example, by centrifugal forces or hydraulically controlled cylinder pivoted out of the gap, possibly fixed and only when starting again is released.
  • Cycle sequence of the jerk crank motor with rotating crank (double crank)
  • In the position of the piston in the Z-OT are in such a construction of the connecting rod, the coupling and the engine crank sequentially switched on a line on the crank shaft. Immediately after the combustion process, when the piston is under the maximum pressure of the working gases and exerting the maximum force in the cycle on its connecting rod, the coupled crankpins of the crankshafts, first by inertial forces, move in different directions from this line. As a result of the fact that the links of the cascade (connecting rod, coupling and engine crank), all at the same time tipping apart in the reciprocal directions, the piston on the jerk crank mechanism can sink extremely fast already at its TDC and with reduced resistance. This allows the working gases to jerk the piston immediately after combustion, thereby increasing the engine's expansion speed many times. At the same time in the immediate vicinity of the OT also the Wirkhebelarm the engine crank is growing rapidly, and is the big in this area connecting rod power here earlier available. This results in dynamic torque growth, which now accelerates the engine crankshaft earlier and more vigorously, and further shortens the duration of expansion of the working gases and further reduces heat losses. Fig. 5 shows the unfolding curves of the active lever arms of the crank of the motor shaft, which are the connecting rod force in the jerk crank drive with different distances available, compared to the development of the curve of the acting lever arm of the crank of the motor shaft of the HM before. In the drawing 6 the set of curves of the torques in the working cycle of the same jerk crank mechanisms with the torque curve of the HM, the maximum value is set as 1, compared. In 5 and 6 the curves of the active lever arms and the curves of the torque of the crank mechanisms are given the same parameters with the same item numbers:
    1. 1. Jerk-crank drive curves with a distance of 1.25 kl;
    2. 2. Jerk-crank drive curves with a distance of 1.375 kl;
    3. 3. Jerk-crank drive curves with a distance of 1.5 kl;
    4. 4. Jerk-crank drive curves with a distance of 1.75 kl;
    5. 5. Jerk-crank drive curves with distance 2,0 kl;
    6. 6. HM curves.
  • It is plausible that the effective lever arm of the connecting rod has its maximum value on the crank when it is at right angles to the crank of the cylinder shaft, which depends exclusively on the connecting rod ratio. This applies both to the motor shaft of an HM and to the cylinder shaft of each jerk crank drive presented here. Perpendicular to the connecting rod, it always stands at about 72 ° CA (π / 2 - arctan (1/3) = 71.57 °), because the crank mechanisms to be compared according to the agreement have the same connecting rod ratio of λ = 1: 3. In this position, the connecting rod force in a gasoline engine with λ = ⅓ is only about 20% of their peak value at the TDC (apart from the additional reduction in force by pressure drop of the working gases in the cylinder because of their inevitable cooling). In the case of diesel engines, with a nearly constant pressure process, the pressure value can be higher when reaching the vertical position of the connecting rod to the crank, in contrast to the equalization process in the gasoline engine, which depends on the injection method of a specific diesel engine.
  • Due to the constant and constant phase shift between the peak value of the connecting rod force and the peak value of the working lever crank arm whose values were each set to 1, the product of these parameters (the peak torque) reaches a maximum of only 0.29 at both the motor shaft conventional gasoline engine as well as the cylinder crankshaft of each jerk crank-engine gasoline engine at 24 ° KW. However, the position of the ENGINE SHAFT of a jerk crank engine at the peak torque in the power stroke depends on the parameters of the concrete design, because at the engine shaft of a jerk crank engine, the value of the instantaneous torque of the crankshaft with the current transmission ratio between the waves multiply , This gear ratio has its greatest value at OT (i max > 1) and the smallest at UT (i min = 1 / i max ), and thus the torque is amplified or attenuated. The peak value of the torque at the engine crankshaft of a jerk Crank drive is therefore usually greater than 1, which is not possible with the HM. This happens with the jerk crank mechanism when the instantaneous transmission ratio of the crank mechanism is greater than the reciprocal of the product of the instantaneous force and the momentary action lever. For example, in the jerk crank 1.25 / 3, the i max = 9 and therefore the relative peak torque 8.65 with the engine crankshaft position of 2.56 °. The degree of fluctuation of the gear ratio as mentioned above depends strongly on the distance, and is larger, the smaller it is. At the same time, however, depends on this distance, the growth rate of the effective lever arm for the piston force: the later the maximum Wirkhebelarm is achieved, the smaller the gas pressure and a large part of this pressure drop is due to the heat loss, which is approximately proportional to the duration of the contact Working gases with the cylinder walls are. The same cylinder charge has the same energy, so it would be somewhat irrelevant in what amount of time this would be converted into mechanical work and transmitted to the motor shaft, if there were no heat loss. In a real engine, however, because one can not avoid the cooling of the working gases, at least the expansion process must be carried out as quickly as possible in order to reduce the heat losses by shortening the duration of contact of the gases with the walls. Exactly this takes place in a jerk-crank engine.
  • Table 6 and the curves of the torque in 5 as well as lever arms of the connecting rod force in 6 show the differences between a crank mechanism of a HM and the jerk crank mechanisms:
    • ✓ In the case of a jerk crank mechanism, the maximum effective lever arm on the engine crankshaft for the piston force (columns 4 and 5), depending on its parameters, is approx. 2 to 5.5 times faster than the basic HM;
    • ✓ The maximum torque in the power stroke (columns 9 and 10) is about 3 to 9 times higher than the torque of the HM;
    • ✓ The maximum torque in the power stroke (columns 11 and 12) is reached about 3 to 9 times faster.
    • 7. Jerk crank mechanism with non-rotatable crankshafts (rocker arms).
  • The return crankshaft is only able to run in the jerk crank mechanism if the following unavoidable conditions exist:
    1. 1. equality of the distance with the length of the coupling;
    2. 2. Equality of crankshaft crankshaft radii.
    Such a construction has from both sides of the crank on the line of the crankshaft axis an EOP at which their direction of rotation opposite to the direction of rotation of the motor shaft when starting by divergers set, or must be monitored (s. 1 - 1, a and b).
  • If the second condition is not complied with in the jerk crank mechanism when the first condition is met, only one crankshaft can run in the full circle, namely the one with the smaller radius of curvature, the other oscillates only at a certain angle about its axis, and the coupling mechanism is transformed a double crank in a rocker arm. But because in a mechanism that is to serve as a crank mechanism for an engine, the motor shaft must always rotate in the full circle, only the Kröpfungsradius the crank handle in a jerk crank mechanism may be greater than the crank radius of the engine crank (s. 17, a and b).
  • Rocker arms with oscillation angles between 0 ° and 180 °
  • If the distance of the length of the coupling equal and the length of the crank handle is greater than the length of the engine crank, a double crank, as mentioned, turns into a rocker arm. When the engine crank is rotating, the return crank oscillates at an angle 0 ° <α <180 ° and the oscillation sector lies only on one side of the crank drive shaft, see 17 , Such a mechanism allows the motor to operate in two different modes in the same cycle. Twice per revolution of the engine crankshaft takes place here a change between jerky and flowing movements at both dead centers of the piston. When turning the motor shaft clockwise and oscillating the motor shaft Connecting rod coupling joint right of the crank drive shaft (s. 17, a ) or in the counterclockwise rotation of the motor shaft and oscillation of the joint to the left of the crank shaft, the piston moves smoothly from the TDC (TDC2) to the TDC (BDC2) and from there abruptly to the TDC (TDC2), like the curve 1 of the diagram of the 18 shows what is not recommended for a crank mechanism of a reciprocating engine. The sine curve (curve 3) serves as a guide. In the direction of rotation of the motor shaft counterclockwise and the oscillation of the connecting rod coupling joint to the right of the crankshaft axis s. 17, b or clockwise rotation of the motor shaft and oscillation of the connecting rod link joint left of the crank shaft axis, the piston moves abruptly from the TDC (TDC1) to the UT (BDC1) and from there in the reverse direction to the TDC (TDC1), such as the curve 2 of the diagram of 18 illustrated. This provides the ability to arrange the strokes of an engine as follows:
    1. a. Sucking jerkily,
    2. b. Compacting,
    3. c. Works jerkily,
    4. d. Ejecting fluently.
    One of the advantages of this rocker arm over other jerk-crank train designs is that it does not have EOPs with its jerk crank mechanism equipped with it, its motion is clearly defined by the direction of rotation of the motor shaft and can not change randomly, so it does not need any devices for directional determination of it kinematic chain links. When making the crank drive, however, it must be noted that the coupling is installed from the side of the crankshaft axis at which it reaches the UT at a smaller angle when starting from the Z-OT with the intended direction of rotation of the motor shaft and a jerk of the piston from the TDC to the UT and not vice versa. Compared to the jerk crank mechanism with revolving crank (double crank), the compression is carried out here more fluently, the duration of the suction decreases, but it remains intense, the passive Ausströmungsphase the ejection shortens slightly and its active phase is comparable to one of the HM. The connecting rod only ever wobbles from one side of the cylinder axis and the piston stroke is less than twice the crank crankshaft crank length. The smaller the difference of the lengths of the cranks the less the deviation from the stroke value 2 kl. For larger differences in the radii of curvature, it may be rational to pitch, offset, or simultaneously tilt and offset the cylinder to the side of the cylinder crank oscillation sector (see 23 ). Thus, a smaller and bilateral uniform Pleuelneigung to the cylinder axis can be achieved and the connecting rod can be shortened. The resulting loss of the quality of the movement curve of the piston is not significant and can be easily corrected if necessary with the change in the distance. A skewed or staggered cylinder axis design also reduces the normal forces resulting from strong unilateral cylinder-axis leaning between the cylinder sleeve and piston skirt surface in a plunger construction and sliding bearing wall in a cross-head design, and distributes them more uniformly to both sides of the guide surfaces. Thus, a reduced and more uniform wear of these surfaces is achieved, which helps to increase the life of the engine.
  • Rocker arms with oscillation angles between 0 and 360 °
  • Suitable crank mechanisms for a reciprocating engine may also be rocker arms whose oscillation angle value is between 0 and full angle (0 ° <α <360 °). However, such a construction as a crank mechanism works under some conditions. If we denote the difference between the lengths of the jerk and the motor crank as Δ (Δ = R1 - R), then only one of the following conditions must be true for it:
    • either Ruck crank > Motor crank and ( paddock + Δ ) = ( distance + motor crank )
      Figure DE102017003146B3_0003
    • or Ruck crank > Motor crank and ( paddock - Δ ) = ( distance + motor crank ) ,
      Figure DE102017003146B3_0004
  • On the basis of such a linkage (s. 15 and 16 ) creates a crank mechanism with a to the crankshaft axis symmetrically to the crank shaft axis in the bow swinging connecting rod coupling joint. These rocker arms also have the property of operating the engine in two different modes in the same cycle. Again, the piston movement is changed from the conventional flowing mode to the new jerky mode during one cycle and vice versa. In contrast to the previously described rocker arm in Section 7.1, however, a change in the type of movement of the piston takes place here only after each complete revolution of the engine crankshaft. Depending on the parameters of such a rocker arm, this happens only in their upper or only in their lower dead centers, so these are Swingarms only suitable for four-stroke engines. Only one of two EOPs of a double crank remains, from which the direction of rotation of the crank handle must be coordinated with the direction of rotation of the motor shaft when initiating the motor startup. The other EOP splits and forms an OT (TDC1 and TDC2 in each side of the vibration sector) 16 ) or a UT (BDC1, BDC2 in 15 ). At these dead points, the direction of movement of the crank arm definitely changes and need not be monitored. The solid broad line ( 15 ) shows the movement curve (1) of the piston, which arises when starting from the EOP (TDC1) with the same rotation of the crankshaft, dashed wide is the movement curve (2) of the piston shown when starting from the EOP (TDC1) the opposite direction of rotation the cranks is selected. The course of the curves 1 and 2 is identical, it has only a phase shift of 360 °. Solid line (3) represents a sinusoidal curve as a guide only.
  • When constructing, however, one must decide at which point the vibration-free circular cut-out of the crank handle should be: If the coupling length is greater than the distance selected by the difference between the crank lengths, the vibration-free circular cut-out of the crank crank between the crankshafts (s. 15 ) and the connecting rod pivot joint oscillates between the piston and the crankshaft. The curve of the piston movement is symmetrical about the vertical axis by its TDC. The duration of the movement from BDC1 to TDC2 and back to BDC2 is less than one revolution of the engine crankshaft. The piston alternately crosses the OT (TDC2) quickly with a jerky double stroke, while the OT (TDC1) passes slowly with a conventional flowing double stroke, so its motion curve is sharply pointed at the TDC (OT) and almost sinusoidal at the TDC (TDC1) , Here the following handling of the cycle would be expedient:
    • Sucking in,
    • Compacting jerkily,
    • Works jerkily,
    • Ejecting fluently.
    Its UTs (BDC) happen here the piston with minimal speed. They both move closer to the OT (TDC2), which is chosen to be the Z-OT, so the compression and power strokes are shortened while simultaneously increasing the suction and ejection, which is beneficial for all cycles of the cycle.
  • If the coupling is selected by the crank difference shorter the distance, the connecting rod coupling joint pierces the EOP between the cranks and now the vibration-free circular cutout of the crank lever lies between the piston and the return crankshaft (s. 16 ) and the connecting rod coupling swings between the crankshafts. The piston movement curve of such a crank mechanism is symmetrical here only about the vertical axis through its UTs (BDC). The curve 1 (solid line) shows the movement of the piston when starting from the BDC1, which simultaneously represents an EOP, with opposite direction of rotation of the crankshaft, curve 2 (dashed wide line) represents the movement curve of the piston when starting from the BDC1 the same direction of rotation of the cranks is selected.
  • When starting from the EOP in the opposite direction of rotation of the crankshafts (curve 1), the piston from the UT (BDC1) runs after a delay of approx. ¾ revolution of the motor shaft with a jerky stroke to the TDC (TDC2), passing the UT smoothly with conventional double stroke (BDC2) and reaches the OT (TDC1), which he leaves abruptly and comes back to the UT (BDC1). Here it would be useful to assign the clocks as follows to the jerk crank drive motor:
    • Sucking in,
    • Compacting,
    • Works jerkily,
    • Ejection jerky.
    Thus, the charge cycle is shortened: it takes place less than in a revolution of the engine crankshaft, the power stroke with the ejection last longer than one turn accordingly. This curve could have some disadvantages compared to the OT symmetric curve ( 15 ), in which the piston remains within the range of its 5% stroke from TDC only about one-tenth of the equivalent length of stay of a HM. The movement curve of the piston is slightly flattened here above, but he still keeps in the range of its 5 percent lift from the TDC compared to the HM but only about 65%, which is lean Mixtures with slower combustion could possibly be beneficial. The ejection becomes longer and thus enormously easy, the suction loses some of its intensity compared to the curve 1 in 15 , but still takes much longer of a HM, the compression curve is identical to that of a HM.
  • The greater the difference of the crank lengths in the jerk crank engine, the smaller the oscillation angle of the crankshaft at the same distance, regardless of the position of the vibration sector. If the impassable circular section of the cylinder crank is between the return crankshaft and the piston, the stroke will be less than two crank lengths of the motor shaft (s. 16 ). The piston moves jerkily from its Z-TDC (TDC1) to BDC1, but its entire stroke takes longer than half a revolution of the motor shaft and it also takes more time for its intended stroke of 95 percent stroke, because its motion curve is not as steep as that of the constructions with rotating crankshafts. The stroke from UT (BDC1) to its GW-TDC (TDC2) also happens jerkily with a delay, but lasts longer than half a revolution of the motor shaft.
  • If the impassable circular section of the crankshaft between the engine and Ruckkurbelwelle, the stroke is greater than two engine crank lengths (s. 15 ). The full stroke from the Z-OT (TDC2) to its UT2 (BDC2) is jerky and less than half a revolution of the engine crankshaft. The power stroke can therefore be carried out very quickly and economically. The stroke from the UT2 (BDC2) to the GW-TDC (TDC1) is conventional and requires more than half a revolution of the engine crankshaft, resulting in a relieved discharge. The suction is fluent, but is prolonged, but loses in intensity compared to the construction with a rotating crank.
  • Advantages of a jerk crank motor with rotating crankshaft in the cycle compared to the base HM.
  • The advantages in the cycle of a reciprocating engine with a jerk-crank-1.25 / 3 with rotating jerk crankshaft with a length of coupling of 1.25 kl, which is equal to the distance and has a connecting rod ratio of 1/3, compared to the cycle of Basic HMs are cyclically described for clarity (s. 1, a )):
  • 1. SUCTION from 0 ° CA to 317 ° CA, duration 317 ° (without feedforward control). When the intake valve is open, the piston returns 95% of its lift from the TDC at 43 ° C. This now achieves much earlier and much lower negative pressure in the cylinder space than with a HM, which leads to forcing the suction, and to intensify the mixture Verwirung, which causes a more perfect combustion. The intake and compression take together, as with a HM, a complete revolution of the engine crankshaft, of which is used for compression in the jerk-crank drive engine only 43 °, the rest of 317 ° (= 360-43) is completely the suction for 150% of the induction time of a basic HM (without pilot control). The nearly motionless stay of the piston of a jerk-crank drive engine during good ¾ revolutions of the motor shaft in the area of the UT in its retardation zone ensures in its upper speed range for a much better cylinder filling and increases the delivery of the engine in general. In summary, the Ruck crank drive engine has a rational torque (as measured by HM criteria) available in a 1.5 times wider band of RPMs than an HM, and its better delivery efficiency, thanks to the more intense cylinder build-up higher displacement specific power and also an absolutely higher maximum power.
  • 2. COMPRESSING from 317 ° CA to 360 ° CA, duration 43 °. The approximately three times faster compression of the mixture ensures less losses of the resulting compression heat. Now, expanding the working gases in the power stroke recovers a greater portion of the energy invested in the compression, as the gases heated by faster compression run less cool.
  • 3. WORKING HAND from 360 ° KW to 403 ° KW, duration 43 °. After combustion of the fuel and reaching the maximum pressure in the combustion chamber at the TDC, the working gases drive the piston towards UT. In short eighth turn of the motor shaft, he puts back 95% of its stroke. The existing heat of the working gases is now given away by the almost four times shorter expansion period less to the walls of the combustion chamber. Due to this situation, the temperature and thus also the pressure of the working gases in the expansion space of the jerk-crank drive motor is continuously higher relative to the pressure of an HM at the same piston position during the entire working cycle. As a result, a greater force permanently acts on the piston in the power stroke, which generates a higher torque on the motor shaft and converts a greater part of the thermal energy into mechanical work. In addition, now grows by the rapid inclination of the coupling of the Wirkhebelarm the engine crankshaft, on which the larger piston force works, also faster. Chart in 6 shows that compared to the base HM (curve 6), the jerk crank mechanism-1.25 / 3 (curve 1) reaches its maximum lever arm 6 times earlier (12 ° CA vs. 72 ° CA).
  • Because close to the TDC in the area of the maximum gas pressure in the cylinder of the large force of the piston, here also a larger active lever arm of the engine crank is available in good time, the force can be implemented earlier, and a higher torque at the motor shaft is reached earlier. The diagram in 5 shows the curves of the relative torque development of the same engines compared to the base HM. For simplicity's sake, the maximum torque of the HM has been assigned the value 1 by analogy. It is obvious that in comparison with an HM (curve 6) the jerk crank 1.25 / 3 (curve 1) has a 8.6 times greater peak torque, which he also 9 times (= 24 ° / 2,6 °) reached faster. As a result, the shaft of the motor is accelerated earlier and more powerfully in the working stroke than in a HM and reaches a higher speed at the end of the working cycle. In order to reduce the heat losses resulting from the shortening of the rotation angle range of the gas expansion by the more suitable function of the jerk crank mechanism, in addition a reduction of heat losses by shortening the time in which the already shortened rotation angle range is traversed, added.
  • After returning a 95 percent stroke at 403 ° CA, the power stroke is virtually completed and the exhaust valve opens. In the next ⅜ revolutions of the motor shaft (403 ° CA to 540 ° CA), the piston only slows down the small remaining 5% part of its stroke to UT. For the jerk-crank-1.25 / 3 this phase of the stroke is longer by 92 ° (= 495 - 403) than in a HM, where the exhaust valve usually opens at 495 ° KW, although its Kobenweg only 89.58% of the stroke has reached. In the case of a HM, such an early termination of the power stroke is made to facilitate the discharge of the exhaust gases, resulting in an expansion loss of 9.38%. This is also justified in the case of a HM, since the economy of the engine does not depend solely on the profitability of the power stroke, but on the economy of the cycle as a whole. In the case of the jerk-crank drive engine, the exhaust valve may be opened later or not at all before the UT is reached, if an offset of the cylinder axes is made (see curve 8 in FIG 3 ) and yet its ejection extends significantly and is thus significantly relieved. By aborting the working stroke before reaching UT, the jerk-crank drive motor will experience an expansion loss of only 4.57% in this case. The HM's expansion loss is 4.81% (= 9.38 - 4.57) greater than the loss-on-loss of the jerk crank motor, which corresponds to a loss of 0.85% in absolute efficiency, which is spared the jerk crank drive engine ,
  • 4. EXTERIORS from approx. 403 ° KW to 720 KW °, duration: 317 °. Ejection in the jerk-crank drive engine is basically in two phases: the passive outflow phase (the first ¾ revolutions of the crankshaft) and the active phase of the exhaust displacement, the actual ejection (the last 1/8 turn of the crankshaft before TDC). In its first passive outflow phase, the piston from the engine crankshaft position of 403 ° CA, when the exhaust valve is opened upon reaching its 95 percent lift, slows the remaining 5% of its stroke back to UT in about ⅜ revolutions of the crankshaft. In the UT (540 ° CA), he changes direction and goes through in the next ⅜ revolutions of the crankshaft again slowly the first 5% of his stroke to TDC. During this phase, during which the piston stops ¾ revolutions of the motor shaft almost motionless near the BDC and the exhaust valve is open, the exhaust gases are discharged from the cylinder. In the second active phase of the ejection, the remaining gases now exhausted are expelled from the cylinder by the piston. The total ejection of the jerk crank motor is 141% of the duration of an HM. However, because compared to the HM results in over three and a half times longer exhaust gas outflow phase (274 ° to 76 °), the residual pressure of the exhaust gases in the cylinder before its second active phase is much lower and they can be displaced with minimal energy expenditure from the cylinder chamber. This also relieves this cycle and additionally contributes to increasing the efficiency in the cycle.
  • Valve timing
  • Apart from the static basic camshaft settings at their installation, which must always be done to adapt the valve timing to a real design in the jerk crank engine, dynamic variable valve timing and variable load behavior also has a positive effect on increasing its efficiency impact. As needed, this variable valve timing control can be used to increase performance, torque gain, or fuel economy during operation. In this aspect, the potential for improvement of the more flexible jerk crank-drive engine is much greater than that of the rigid HM in the present state of the art. First, the jerky pushing out of the residual exhaust gases in the final phase of the exhaust stroke triggers the jerk-crank drive engine directly on the TDC last still a short increase in pressure of the exhaust gases in the cylinder. This causes an intense and helpful suction effect in the exhaust tract immediately after the piston passes its OT. In combination with the jerky piston movement to the UT, which occurs immediately after ejection with a much lower delay than an HM, because the piston passes the GW-OT much faster in the jerk crank drive engine, it firstly has a positive effect on purging off when the exhaust valve closes with the intake valve open with deceleration to the instantaneous rotational speed deceleration. Secondly, thanks to a strong downwards acceleration of the piston near the TDC and at the same time the rectified jerky submersion of the intake valve disk (s) on opening, the fresh gas from these bodies is entrained mainly in the direction of the cylinder axis. With the remaining remaining at the cylinder head of the combustion chamber exhaust gases, they will not mix significantly in this phase, because exhaust gases in the combustion chamber with imparted them from the piston momentum in the last phase of its upward movement already flow towards exhaust valve opening and additionally used by the suction of the exhaust to the exhaust valve become. This convenient combination of this suction with the faster downward movement of the piston causes cleaner purging than with a HM, whereby the exhaust gases can be removed almost completely and with virtually no flushing losses when the exhaust valve is closed with the appropriate delay.
  • A jerk-crank drive motor is quasi an HM, which makes it possible to execute individual cycles at different speeds of rotation with the smooth circulation of its motor shaft in the cycle, thanks to its periodically fluctuating gear ratio of its crank drive. The piston is accelerated twice and slowed down twice in accordance with the course of the thermodynamic processes in the cylinder. This makes the cycle process much more effective and economical compared to a HM:
    1. a. its intake takes about 41% longer: 317 ° to 225 °;
    2. b. the compression happened 3.44 times faster: in 43 ° against 149 °;
    3. c. the expansion also happens 3,44 times faster: in 43 ° to 149 °;
    4. d. its discharge lasts about 41% longer: 317 ° to 225 °.
    With uniform rotation of the engine shaft of a jerk crank engine with 6,000 rpm, it has a more effective intake capacity than a HM at 4 255 rpm and so moderate output losses, as an engine with a conventional crank mechanism at a speed also from ca 4 255 rpm. At the same time, however, the compression and expansion proceeds as sparingly as with an HM at a speed of approximately 21,000 rpm. In this case, the compression and power strokes, in which the heat losses of the temperature difference and the duration of the contact are proportional, not average, but especially in their most wasteful phases, where the temperature difference is highest, maximally accelerated. Due to the much more intensive sequence of the intake stroke and its extension, better filling, rinsing and delivery levels of the engine are made possible, the ejection is greatly facilitated by its extension. The efficiency and its specific power increase.
  • The calculation of the increase in the efficiency of the jerk crank engine.
  • The total conversion losses of an internal combustion engine are set out in Table 7:
    • they are calculated from the value of the theoretically possible efficiency minus the value of the actual efficiency. A common gasoline engine with a common degree of expansion 10 (compression ratio 10: 1) has a theoretical efficiency of around 0.60: η = 1 - ( V1 / V2 ) 0.4
      Figure DE102017003146B3_0005
    • Line 3: η = 1 - ( 1.10 ) 0.4 = 0 6019
      Figure DE102017003146B3_0006
      The gasoline engine (statistically) achieves a de facto efficiency of
    • Line 4: 0 , 25 to 0 30
      Figure DE102017003146B3_0007
      His summary losses are therefore in the range of 0.30 to 0.35:
    • Line 5: 0 , 6019 - 0 25 = 0 , 3519; 0 , 6019 - 0 30 = 0 3019 ,
      Figure DE102017003146B3_0008
      When subtracting the dissipation losses (friction, pump, radiation losses) of the engine of approx. 5% absolute, which are included in the total losses, the pure heat losses of a gasoline engine are rated at 0.25 to 0.30 absolute:
    • Line 8: 0 , 3519 - 0 , 05 = 0 , 3019; 0 3019 - 0 , 05 = 0 , 2519 ,
      Figure DE102017003146B3_0009
  • A common conventional truck diesel engine with a typical compression ratio of 22: 1 has a theoretical efficiency of about 0.71:
    • Line 18: η = 1 - ( 1.22 ) 0.4 = 0 , 7096 .
      Figure DE102017003146B3_0010
      its actual statistical efficiency is between 0.35 and 0.40. His summary losses are in the range of 0.31 to 0.36:
    • Line 20: 0 , 7096 - 0 35 = 0 , 3596; 0 , 7096 - 0 40 = 0 , 3096 ,
      Figure DE102017003146B3_0011
      After subtracting dissipation losses of 5% absolute, the values of its heat losses are in the range of about 0.26 to about 0.31:
    • Line 23: 0 , 36 - 0 , 05 = 0 , 3096; 0 , 31 - 0 , 05 = 0 2596 ,
      Figure DE102017003146B3_0012
      The dominant representative heat losses in a HM occur in the working stroke and in the last phase of the compression stroke and are approximately proportional to the duration of these cycles. However, with the jerk-crank-drive motor, as these lossy strokes are faster, it is safe to say that heat losses will also decrease by 3.44 times in proportion to this acceleration: 148 , 7 ° / 43 2 ° = 3 , 4421 ,
      Figure DE102017003146B3_0013
      However, considering that the expansion and compression in the vicinity of the TDC, where the temperature difference of the working gases with the cooled cylinder walls is highest, increases disproportionately in the jerk-crank drive engine, the savings are higher. Assuming a uniform rotation of the crankshaft of the internal combustion engine, one can substitute the time t with the aid of a proportionality coefficient C by the rotation angle φ. The temporal change of the individual factors takes into account then for the heat losses within the engine cycle: Q = ω ψ C α S ( φ ) Δ T ( φ ) d φ
      Figure DE102017003146B3_0014
      ω and ψ
      in this case enclose the rotation angle range, which corresponds to the time period in which the heat transfer between the working gas (mixture) and the cylinder walls takes place.
      α
      the heat transfer value between the working gas and the cylinder walls remains constant (independent of pressure),
      S
      the working gas enclosing surface changes with the piston stroke,
      .DELTA.T
      the temperature difference between the working gas at temperature T 1 and the cylinder walls at temperature T 2 changes with the compression or expansion of the working gas.
  • In this way calculated degree of reduction of heat losses reaches namely in the jerk-crank drive engine a value of 3.56, or the heat losses account for only 28.09% of the heat losses of a HM. The share of total saved heat losses in the cycle in the working cycle is just under 89% and in the compression cycle a good 11%. These results are obtained assuming that the temperature of the fresh mixture is 70 ° C and the temperature of the cylinder walls is 100 ° C. If one is interested in values of Q RKM / Q HM on the basis of other temperature parameters, note that the value of Q RKM / Q HM differs by less than 1.5% from the calculated value 28.09% as long as the temperature of the fresh mixture (T m ) and the cylinder walls (T 2 ) are selected within the following limits: T m : from 333K to 373K ( 60 ° C to 100 ° C ) ; T 2 : from 373K to 423K ( 100 ° C to 150 ° C ) ,
    Figure DE102017003146B3_0015

    The saved share of heat losses through acceleration of the working and compression stroke is therefore 71.91%: 1 - 0 , 2809 = 0 , 7191 ,
    Figure DE102017003146B3_0016

    This fraction is converted into mechanical work and contributes to increasing the efficiency of the engine. The increase in efficiency for the Otto-Jerk crank drive engine is thereby at least 18.11% and at most 21.71% absolute: 0 , 7191 × 0 25 = 0 1811 and 0 , 7191 × 0 30 = 0 2171
    Figure DE102017003146B3_0017

    and its absolute efficiency reaches a value of 46.71 to 48.11%:
    Line 9: 0 25 + 0 2171 = 0 , 4671; 0 30 + 0 1811 = 0 , 4811 ,
    Figure DE102017003146B3_0018

    A jerk crank-drive gasoline engine thus increases its efficiency relative to the HM alone by optimizing the working and compression stroke by 60.38% to 86.84%:
    Line 11: 0 , 4671/0 25 = 1 , 8684; + 86 , 84% relative; 0 , 4811/0 30 = 1 , 6038; + 60 , 38% relative ,
    Figure DE102017003146B3_0019

    Its specific fuel consumption drops to a level of 53.52% to 62.35% of the consumption of a HM:
    Line 15: 1 : 1 , 8684 = 53 , 52%; 1 : 1 , 6038 = 62 , 35%
    Figure DE102017003146B3_0020
  • An identical calculation of the heat losses of the jerk crank-drive diesel engine shows that in its cycle the proportion of heat losses to a level of 18.67% to 22.26% of the heat losses of an HM drops completely: 0 2596 × 0 , 7191 = 0 1867 . 0 , 3096 × 0 , 7191 = 0 , 2226 ,
    Figure DE102017003146B3_0021

    Thus, the efficiency of a jerk-crank-diesel engine reaches a value of 57.26 to 58.67% absolute
    Line 24: 0 35 + 0 , 2226 = 0 , 5726; 0 40 + 0 1867 = 0 , 5867 .
    Figure DE102017003146B3_0022

    which corresponds to an increase in its real relative efficiency over the conventional diesel engine from 46.67 to 63.61%:
    Line 26: 0 , 5726/0 , 35 = 1 , 6361 : + 63 , 61% relative; 0 , 5867/0 40 = 1 , 4667: + 46 , 67% relative ,
    Figure DE102017003146B3_0023

    The fuel consumption of a jerk-crank-diesel engine thus drops to a level of 61.12% to 68.18% of the consumption of a conventional diesel engine:
    Line 30: 1 : 1 , 6361 = 61 , 12% 1 : 1 , 4667 = 68 , 18%
    Figure DE102017003146B3_0024
  • Calculation of the power increase of the jerk crank engine.
  • The economical realization of the working and compression stroke, extended and more efficient intake with a more efficient cylinder charge and at the same time relieved ejection of the jerk-crank drive engine increase its torque and increase its efficiency. The increased cycle performance of a jerk crank motor allows it to operate at a lower RPM range than a comparable HM to accomplish the same work. This additionally reduces the mechanical losses and wear of the motor as it requires fewer cycles to perform the same job. On the other hand, a jerk-crank motor achieves a significant increase in performance by qualitatively improving the functions of each of its individual clocks at the same rotational speed. Thus, the somewhat heavier construction of the mechanism of the jerk-crank drive engine is in addition to increased efficiency compared to a significant reduction in the power weight.
  • In terms of performance, it can be argued that increasing the speed of a jerk crank engine by 41% over the RPM of an HM at the point of its peak torque can still provide at least as much torque as an HM because of the actual torque Duration of the suction, which is 41% longer in the cycle, thus does not decrease and the degree of its cylinder filling is maintained (only the duration is taken into account, the increased intensification of the intake of the jerk crank engine remains out of consideration!) , The increase in speed by 41% and thus occurring shortening of the ejection is also compensated by the generally 41% longer ejection of this cycle in the jerk-crank drive motor and therefore remains the same duration of ejection of a HM. These facts allow convinced to confirm that this clock is not difficult and the total cycle losses of a jerk-crank drive motor at elevated 41% rotation speed at least equal to those of a HM remain. On the other hand, the torque of a jerk crank-crank gasoline engine increases by 60% to 87% and a jerk crank-drive diesel from 47% to 64% by reducing the heat losses (see Table 7). Therefore, the output of the jerk crank motor increases at rotational speed corresponding to its maximum torque firstly by increasing the rotational speed and secondly by increasing the torque. This gives a 2.26 to 2.63-fold increase in output of the jerk crank-type gasoline engine or a 2.07 to 2.31-fold increase in output of a jerk crank-drive diesel:
    • Line 14: Jerk crank drive gasoline engine: 1 409 × 1 , 6038 = 2 , 26; 1 409 × 1 , 8684 = 2 , 6 respectively ,
      Figure DE102017003146B3_0025
    • Line 29: Jerk crank diesel engine: 1 409 × 1 , 4667 = 2 , 07; 1 409 × 1 , 6361 = 2 , 31 ,
      Figure DE102017003146B3_0026
  • The presented calculation of the increase of the efficiency and the performance of a jerk crank engine considers only the reduction of the heat losses, which by acceleration of the Working and compression stroke come about, which occur because of the more appropriate function of the jerk crank engine. However, it did not take into account:
    1. 1. That, in addition to the reduction of the heat losses resulting from the shortening of the rotation angle range in which the gas expansion is performed with uniform rotation of the motor shaft, in addition to reducing the heat losses by shortening the time is added, in which the already shortened rotation angle range is traversed added ;
    2. 2. that by increasing the efficiency of the gas exchange, the cylinder charge has risen (higher rinsing, catching and delivery), whereby greater torque and cycle work is done;
    3. 3. that the performance of the cooling system can be lowered because of the reduced heat losses (decreasing internal self-consumption, weight reduction);
    4. 4. That the degree of expansion in the power stroke in the jerk crank engine from 95% to 100% of its geometric stroke amounts to 89.58% in the HM, which increases the efficiency of the Otto-Jerk crank-drive engine by 0.86% to 1 , 60% absolute (from 3.44 to 5.33% rel.) And the diesel jerk crank-drive engine by 0.67% to 1.24% absolute (from 1.91 to 3.74% rel.) improved;
    5. 5. That the sealing losses are minimized because of the enormously shortened duration of the working and compression cycle.
  • Environmental compatibility and emissions.
  • Firstly, the jerk-crank drive engine is more efficient than the conventional internal combustion engines and therefore has smaller specific heat emissions, thus contributing less to direct climate warming. Second, the power-to-weight ratio of the jerk crank engine is estimated to be at least 1/3 higher than that of an HM. This means that 1/3 less metal is needed to produce jerk crankshaft engines with equivalent total power at today's level, so CO2 emissions for engine manufacturing will also be reduced by 1/3 , Third, its gas emissions, which indirectly affect global warming through greenhouse effect, are also lower. The emissions of CO2 are reduced proportionally to the specific fuel consumption of the jerk-crank drive engine and thus have a value: in the jerk crank-engine gasoline engine from 53.52% to 62.35% and at the jerk-crank drive diesel engine from 61.12% to 68.18%
    from the emissions of a HM.
  • Fourth, the pollutant emissions of a jerk-crank drive engine fall, except for the CO2 disproportionately, because:
    1. a. they decrease in proportion to the specific fuel consumption, which is about 38 to 47% lower for gasoline engines and 32 to 39% for diesel engines, as well as CO2 emissions;
    2. b. The conditions for the formation of harmful thermal NO x , which are positively influenced by both the high temperatures and the residence time, deteriorate significantly in the combustion and expansion space of a jerk-crank drive engine, because:
      1. i. the temperature of the working gases, due to faster expansion, because the power stroke in the jerk-crank-drive-engine cycle generally runs about three and a half times faster, decreases faster;
      2. ii. the reaction time with high gas temperatures, at which these pollutants increasingly form, is reduced exponentially fast.
  • According to Zeldovich's thermal "NO mechanism", which describes the oxidation of atmospheric nitrogen, which is also used for the combustion processes of engines, it is known that the formation of thermal NOx from about 1300K is to be expected. The rate of formation increases exponentially at temperatures above 2200K and a sufficient oxygen concentration (λ> 1). Analysis of the temperature curve during the expansion of the working gases in the working cycle of the engines to be compared shows that assuming the starting temperature of the working gases in the combustion chamber at the working cycle of 2273K (2000 ° C), the temperature above the limit of 2200K, at which nitrogen oxides form exponentially Jerk crank engine 1.1 ° and HM 9.9 ° lasts, so it lasts nine minutes longer in the HM. Here, however, the area remains with temperatures above 2200K, where an exponential nitric oxide formation occurs because of the complex calculation and the short phase of this state (about 15% of the duration of the cycle with temperature of the exhaust gases over 1300 ° C) out of consideration, which is disadvantageous for the invention. The evaluation of the total time span of the expansion with the starting temperature of the working gases from 2273K in the working cycle to the lowering of the temperature up to 1300K is carried out as if only a linear nitrogen oxide formation takes place in this period. With the HM, the temperature drops from the beginning of the power stroke to the limit of 1300K with 62.7 ° CA, a jerk-crank drive engine 1.25 / 3 reaches this lower temperature limit already with 7.7 ° KW. For the Jerk Crank Engine 1.25 / 3, the time to cool down to the 1300K level is 8.143 (= 62.7: 7.7) times shorter. This also results in a correspondingly smaller amount of nitrogen oxides than the HM. Taking into account the fuel reduction, in terms of increasing the efficiency, a total reduction in specific emissions of NOx (g / kWh) will be at least 13.13 to 15.36 times in the Otto-Jerk crank drive engine and at least 11.97 to 13.35 times achieved with the Diesel-Jerk crank drive engine: Otto Ruck crank drive engine: from 8.143: 0.6235 = 13.13 to 8.143: 0.5352 = 15.36; Diesel-jerk crank drive motor: from 8.143: 0.6818 = 11.97 to 8.143: 0.6112 = 13.35.
  • If an engine with internal mixture formation is operated at full load, the mixture can easily be maintained at a stoichiometric combustion ratio (λ = 1). In part-load and idling (especially diesel) operating modes, the superfluous air volumes cause only lean mixture to form and produce a large excess of oxygen (λ> 1), favoring unwanted NOx formation. This process can be effectively counteracted by using a portion of the residual exhaust gases without oxygen instead of fresh air for cylinder filling for cylinder filling, so the known exhaust gas recirculation can be used.
  • Emissions of hydrocarbons and carbon monoxide, which today can be oxidized with known exhaust aftertreatment by catalyst almost without problems, are not treated here.
  • Embodiments of the jerk crank drive motor.
  • A jerk-crank motor can be built just as any, as well as a HM: as a single-cylinder, boxer, V, series, X, H, radial engine; designed according to two- and four-stroke procedures; be designed as gasoline, gas or diesel engine; be carried out as a naturally aspirated or charged. However, with the same displacement, a jerk crank motor achieves a much greater peak in torque in the power stroke than a HM. This pulsation of torque causes a higher rotational irregularity of the motor shaft in its cycle, which requires a larger flywheel. To get better smoothness, it is advantageous to build a jerk crank engine as a multi-cylinder engine. First, the rotating mass increases in total and the weight of the flywheel can be reduced thereby. Secondly, the necessary balance of the rotating mass forces can be realized with skillful arrangement of the cylinder for the most part with the mass of the opposing rotating structural parts and only the remaining unbalanced mass needs to be balanced with complementary counterweights, thus a better power to weight is achieved. With the jerk-crank drive engine, the mass balance is more complicated than with the HM because of the double number of cranks with several rotating and oscillating links. But in multi-cylinder boxer engines, the mass forces with optimal arrangement of the cylinder can usually compensate well or completely.
  • In the 2-cylinder boxer engine with two-stroke process, which can be well balanced with well thought-out mirrored arrangement of the cylinder, the working cycles in the opposite cylinders occur simultaneously and can not be distributed evenly in one revolution. As a result, its smoothness is comparable to that of a single-cylinder engine, but it is low in vibration.
  • A smoother running has a 2-cylinder boxer block in the four-stroke process (see 21 , and 22 ). Here, all oscillating mass forces 1st and 2nd order, as well as the two-stroke, completely by symmetrical movements of the pistons, connecting rods and coupling completely, but the nonuniformity of its motor shaft is smaller, because his strokes can be evenly distributed and in each revolution the motor shaft passes only one working cycle. However, such a block still has a circumferential mass moment due to the offset opposite cylinder axes, which can be completely canceled only with counterweights or in an inline boxer engine design by means of a crankshaft with three crankshafts and a tandem spindle, which is expensive. But this problem can also be solved differently: namely, if you change it from two mirrored 2-cylinder boxer blocks to a 4- Cylinder boxer block united (see 24, a and b). Thus, the circumferential mass moments without counterweights, as in an inline boxer engine, completely cancel each other out. Although the working cycles of the opposite cylinders are superimposed, they can be perfectly distributed between the opposite cylinder banks. The shaft journals of the motor shaft are freed from the piston pressure forces by simultaneous ignition in the densest opposite cylinders because they compensate each other. This is possible with the following ignition sequence: (1 + 3) - (2 + 4).
  • The height of a conventional crosshead motor, especially in long-stroke two-stroke engines, is relatively large. If he is provided with a jerk crank drive, its height increases unfavorably in addition. To overcome this problem, an engine may be constructed with a jerk crank drive which is at an angle to its cylinder axis with its axis. As a result, its height increases only slightly or not at all when the angle formed between the axes is δ ≦ 90 °. 13, a ) and b) shows on the basis of a cross-head and a plunger engine, the basic formation of such constructions, when the axis of the cylinder is perpendicular to the axis of the crank mechanism. This modification only changes the design of the crank mechanism, without causing any changes in the motion function. Thus, it differs fundamentally from the constructions, modified by means of inclination of the cylinder axis to the axis of the crank mechanism (s. 13, c ), which gives the movement curve of the piston a strong asymmetry. With sole inclination of the cylinder axis to the axis of the crank mechanism (c) the formed by the connecting rod (3) and the coupling (7) common joint (3-7) is maintained and is guided by the single crank pin of the cylinder crankshaft (6). But if the motor shaft (5) is rotated about the cylinder shaft (6) (a and b), the joint (3-7) (see also 14, a and b) divided into two individual joints: the connecting rod (3) is coupled to the crank pin (6/2) and the coupling (7) to the crank pin (6/3) of the engine crankshaft (6). These two crank pins are at an angle δ to each other, which is equal to the angle between the cylinder and crank drive axles. Such a design of the RKM does not increase its height, but its width, because its motor shaft is not placed under the cylinder shaft, but next to it. In this case, when the engine crankshaft is used in common for a plurality of parallel cylinder banks, the multi-cylinder engine thus formed becomes more compact and its power weight at the same height improves. 14 shows a logical evolution of the construction of the jerk crank engine from a single-cylinder to multi-cylinder engine with common engine crankshaft:
    • Sketch a) represents a single-cylinder construction with a motor shaft (5) lying under the cylinder crank (6), where it is coupled to the joint (3-7), which is guided by the single crank pin of the cylinder shaft (6) in a circle ;
    • Representation b) shows a laterally placed motor shaft (5), which is coupled by means of coupling (7) on the additional crank (6/3) of the cylinder shaft (6), which is offset by angle δ to the crank (6/2). The connecting rod joint (3-6 / 2) is guided by the crank (6/2);
    • Drawing c) discloses a boxer engine design where the cylinder crankshaft (6) has two additional in-plane opposed cranks (6/2 and 6/4) for the connecting rods (3/1 and 3/2) and one vertical to this plane aligned crank (6/3) to which the engine crankshaft (5) is coupled by means of the coupling (7) required;
    • Drawing d) shows a four-cylinder H engine, which was formed of two parallel to engine crankshaft (5) coupled units, exactly as in c). The two cylinder crankshafts (6 and 9) of each pair of pistons are coupled with their cranks (6/3 and 9/3) by coupling (7/1 and 7/2) to the intermediate engine crankshaft (5).
  • Such a four-cylinder unit is almost perfectly balanced (except for the circumferential mass momentum due to displacement of the cylinders along its cylinder shafts) and allows for uniform distribution of power strokes in the cycle and gas exchange to the opposite cylinder banks. If desired, the increase in the number of cylinders along the engine crankshaft axis could be continued by connecting another four-cylinder unit to an extended engine crankshaft.
  • 8th shows a two-cylinder boxer block with side-mounted engine crankshaft (5), to which a common crank handle (6) for two cylinders (1) is coupled. However, such a cylinder crank must also be dimensioned for compressive forces and torques, which it is exposed to here in addition to the tensile forces.
  • 9 shows a 4-cylinder jerk crank engine consisting of 2 two-cylinder Boxer blocks with gear synchronization of the crankshaft of each block. If instead of two-cylinder Boxer blocks four-cylinder blocks, as in 24, b and common cylinder banks, which can be arbitrarily extended along the axis of the engine crankshafts, creates a fully balanced multi-cylinder engine.
  • The radii of the coupled together cranks of the waves in 13 and FIG. 14 may be any of the radii of the cranks to which the pistons are coupled to vary the distance between the shafts of the crankshaft journals, as needed, to the design. While maintaining the same ratio of the length of the coupling to the crank lengths, the gear ratio of the linkage remains in the same range and also the piston stroke does not change. If the distance between the engine and cylinder shaft is lengthened, the coupling length and the radii of the coupled crankses need only be adapted to this shaft spacing. For example, in order to reduce the height of the jerk-crank drive motor (s. 13, b )) has its engine crankshaft (5) rotated by 90 ° about the cylinder crankshaft axis clockwise and placed on the left side of her. It had to be in order to accommodate them with necessary distance from the cylinder to accommodate, the original distance of 1.5 kl increased by 35%. In each case both the radii of the crank of the motor shaft (5) and the crank (6-3) of the cylinder shaft (6) and the length of the coupling (7), by means of which they are coupled, have been increased by 35%. In this way, the height of the jerk crank motor was reduced without changing the piston stroke and the original function of the jerk crank mechanism.
  • Opposite piston jerk crank drive motor.
  • 25 shows an exemplary embodiment of a two-stroke piston engine with two parallel cylinders.
  • The displaceable cylinder (1/1) piston (2/1) is coupled by connecting rod (3/1) with the crank pin of the crank arm (6/1) embedded in the bearings (15/1) of the housing.
  • The displaceable in the cylinder (1/2) piston (2/3) is coupled by connecting rod (3/3) with the crank pin of the in the bearings (15/3) of the housing Ruckkurbel (6/3).
  • The engine crankshaft (5/1), which lies under the crank (6 / 1a) (not shown), forms with the engine crankshaft (5/3) (see view CC) by means of the intermediate crank (30/1) of the parallel crank mechanism (30) an inseparable unit that seats in bearings (14/1 and 14/3) in the housing and is located under the cylinder cranks on the left side of the engine.
  • The coupling (7/1) rotatably connects the additional crank (6 / 1a) of the crank (6/1), which is offset by 90 ° against its Pleuelhubzapfen, with the engine crankshaft (5/1), which under the cylinder crankshaft (6/1). 1), analogously, as the coupling (7/2) shown in section AA, the Zusatzkröpfung (6 / 2a) of the crank (6/2), which is offset by 90 ° against the Pleuelhubzapfen the cylinder crankshaft (6/2), coupled with the engine crankshaft (5/2).
  • View C-C shows: the coupling (7/3) connects the Zusatzkröpfung (6 / 3a) of the crank (6/3), which is offset by 90 ° to its Pleuelhubzapfen with the crankpin of the engine crankshaft (5/3).
  • Analogously, the right in the cylinder (1/1) displaceable piston (2/2) by connecting rod (3/2) with the crank pin in the bearings (15/2) of the housing bedded return crank (6/2) coupled.
  • The displaceable in the cylinder (1/2) piston (2/4) is coupled by connecting rod (3/4) with the crank pin in the bearings (15/4) of the housing Ruckkurbel (6/4).
  • The engine crankshaft (5/2) forms with the engine crankshaft (5/4) by means of the intermediate crank of the crank (30/2) of the parallel crank mechanism (30) an inseparable unit in bearings (14/2 and 14/4) in the Housing embeds and lies under the cylinder cranks on the right side of the engine.
  • The coupling (7/4) connects the Zusatzkröpfung (6 / 4a) of the crank (6/4), which is offset by 90 ° against its Pleuelhubzapfen with the engine crankshaft (5/4), s. 25 , View BB and AA.
  • The rotation of the common left (5/1 & 5/3) and the common right motor shafts (5/2 & 5/4) is made with the push rods (31/1 & 31/2) from the cranks (30/1 & 30/2) in the parallel crank gear (30) brought together and passed to the motor shaft (5/0).
    In the parallel crank gear (30) of the crank pin of the push rod (31/2) against the crank pin of the push rod (31/1) to overcome the dead center offset by 90 ° and all three waves have the same direction of rotation.
  • The working stroke of the cylinder (1/1) is offset by the working stroke of the cylinder (1/2) by 180 ° in order to minimize the rotational irregularity on the motor shaft.
  • The two-stroke piston engine with its high power density benefits particularly from the equipment with a jerk crank mechanism:
    1. 1. Its total heat losses, which are already reduced by half due to the lack of cylinder head by practically halved combustion chamber surface, are now quartered by the return crank by shortening the duration of the heat exchange with the cylinder walls. This avoids approximately ⅞ (= 1 - (½ × ¼)) of its current heat losses, and its level drops to about ⅛ that of a HM. Thus, the difference of the practical and theoretical efficiency of an internal combustion engine with such a combination of mechanisms is as low as ever before;
    2. 2. The problem of the life of the piston engine in view of the enormous thermal load due to the insufficient cooling of the outlet side, in particular the top edge of the Auslasskolbens (2/2 and 2/4) and the highly thermally stressed exhaust slots of the cylinder (22/1 and 22/2 ) falls completely off when equipped with a jerk crank drive. With the jerk-crank mechanism, thanks to its typical retardation zone, the pistons now keep more than ¾ of the engine crankshaft revolution close to the UT in the area below 5% of the piston stroke in this problem zone and the duration of the purge increases by a multiple (approx. 7 times!). The enormously prolonged purging process, which is usually a major disadvantage due to fuel losses (flushing losses) when external mixture formation is employed, proves to be quite useful when using internal mixture formation. In the modern direct-injection gasoline and diesel engine, the superfluous amounts of air that flow through the cylinder and can not increase the consumption of fuel, are not harmful. On the contrary, such effective fresh airflow, after pushing out the hot exhaust gases, is needed as much needed cooling for the piston and cylinder. Just this effect was missing the current piston engine, because the flushing of a two-stroke engine with a conventional crank mechanism with approximately sinusoidal motion curve of the piston is extremely short (about 1/5 of the intake stroke of a conventional four-stroke engine at the same rotational speed). This limited period of time is only sufficient for satisfactory flushing when using a positive pressure blower and is not sufficient to provide effective cooling. In 2-stroke engines with a jerk crank mechanism, however, an effective cooling effect can be achieved even without a blower, simply by its enormously extended flushing. For large supercharged engines, the purge fan may well perform this additional effective cooling function. If, for example, a diesel engine is running at full load, the mixture can easily be kept at a stoichiometric combustion ratio (λ = 1). At partial or idle speed, the unnecessary superfluous air volumes, unlike a gasoline engine, cause only a lean mixture to form and excess oxygen (λ> 1) is produced, favoring unwanted NO x formation. But also the thermal load of the engine is lower and you can without risk counteract this process drastic by the irrigation - at the same time the cooling - is insulated without risk. It is expedient to use the purge instead of fresh air partly with the exhaust gases without oxygen for cylinder filling, which is known as exhaust gas recirculation. With large engines, the predominant operating condition of which is full load, it makes sense to automatically control the flow rate of the fresh air with a fan by means of sensory monitoring of the temperature of the exhaust air in order firstly not to overheat the engine and secondly not to unnecessarily increase the own energy consumption through excessive air supply ;
    3. 3. In the twin-piston engines with two crankshafts the cascade of gears is replaced by the jerk-crank mechanism with the parallel crank gear and needs no additional space along the cylinder axis. The expansion along the crankshaft increases only slightly;
    4. 4. By placing multiple cylinders in parallel, with each two cylinder crankshafts at the cylinder ends, which are coupled with an intermediate engine crankshaft, this design can be operated, allowing a low power-to-weight ratio;
    5. 5. This synthesis produces approximately an ideal mass balance for both mechanisms.
  • Additional optimization options for the piston movement of a jerk-crank drive motor.
  • Primarily, the degree of non-uniformity of the piston speed in the jerk-crank motor depends on the ratio of the length of the coupling to the radius of curvature of the motor shaft, secondarily on the connecting rod ratio. By way of illustration, the recessed piston travel is shown in Table 4 with the motor shaft 30 ° CA in the case of various engine types. As you can see, the piston paths differ strikingly: the piston of a jerk-crank drive motor lays, depending on the length of its coupling and the connecting rod length, already at the position of the engine crankshaft of 30 ° KW about eight or ten times the distance one HM back. In a KPM at the same crank position, the piston passes only about 50% greater path than a HM. In Table 8, the characteristic times of the work cycle, piston and expansion rates of the different crank gears are compared, and other relevant values of each engine type relative to the values of the base HM are indicated. Compared to the basic HM, the KPM work cycle has a connecting rod length of 1.0642 and 1.0154 kl (corresponds to a maximum connecting rod inclination to the cylinder axis of 70 ° or 80 °) by 37.0 or 57.4% faster instead. In this case, a comparable jerk-crank drive motor with a length of the coupling of 1.5 to 1.25 kl realizes its complete power cycle by 2.1 to 3.9 times faster than the HM. In contrast to the KPM, however, the piston of a jerk-crank-drive engine accelerates significantly more directly here after the TDC. See the piston speed ratios of the jerk-crank motor in relation to the HM: with a 3.5-fold faster stroke, the maximum piston speed is 6 to 7 times higher and 12 times earlier.
  • A more precise adaptation of the piston movement curves to specific thermodynamic processes, running in the cycle of a specific internal combustion engine, can be achieved in addition to the highest possible efficiency, by means of displacement and (or) inclination of the cylinder axis to the crank pin axis. Thus, e.g. the OT-symmetric curve of a jerk-crank-drive-engine with revolving crank-crank gives an asymmetry, which additionally increases the speed of expansion and facilitates the ejection.
  • In constructions with a correspondingly offset or inclined cylinder axis, the working cycle need not be terminated before the UT is reached. The double stroke of the piston from the Z-OT to the GW-OT, which usually takes a complete revolution of the motor shaft, is distributed asymmetrically around the UT here. The working stroke becomes shorter and the saved time at it is allocated to the ejection. See eg 3 , Curve 8 of a jerk crank-1.25 with cylinder offset by a crank length from the crank axle: here, the distance between his Z-OT and UT is only 48 ° KW. Therefore, in such a construction, no expansion losses occur because the exhaust valve is opened only at the BDC, and yet, it is possible to perform economical exhausting with a duration of 312 °.
  • It is significant that the readjustment in the jerk-crank-train constructions is much more effective than in the case of a HM: so an offset of the axis of the cylinder against the axis of the crank drive brings about a crank length for the jerk-crank-drive-engine-1,25 / 3 shortening the working stroke (end at the 95 percent stroke (0.95s)) by 44% (decreases from 43 ° to 24 °), with a HM the power stroke (0.95s) shortens only 9% for the same displacement ( drops from 148.7 ° to 135.6 °). Curves of some such constructions are an example in the family of curves 3 shown. 22 shows the positions of the members of a 2-cylinder boxer with cylinder displacement in the cycle at the beginning of each cycle during one revolution of the engine crankshaft: upper cylinder during intake and compression, lower during working and discharging. 21 shows for comparison the positions of the members of this engine in the cycle without cylinder displacement.
  • If desired, a jerk crank mechanism can be designed with a variety of piston movement functions. Crankshaft drive, in which the cylinder axis is in line with the crankshaft axis, and has a distance which is equal to its coupling length, at the same radii of the coupled crankshafts crankshafts a rotatable cylinder crankshaft (centric crank on a double crank). Rotating the cylinder crankshaft in such a mechanism in the opposite direction to the engine crankshaft rotation, it is suitable for a crank mechanism and allows the execution of a jerky piston movement whose degree of nonuniformity is in inverse proportion to the distance. The curve of the piston movement that produces such a mechanism is symmetrical and its mirror axis is on the crankshaft axis. In the same direction of rotation of the cylinder and engine crankshaft, the movement function of the piston does not differ from that of a HM, regardless of the distance (but at D = K!). Such a function alone is useless for a reciprocating engine, because it merely copies the function of a conventional crank drive, but it can also be usefully used in combination with other modes of motion. For example, if desired, a jerk crank drive can be designed so that the piston in the cycle fluent in one bar and jerky in the other. Such features have mechanisms of rocker arms (see section 7, a and b), which are shown in drawings 15 . 16 . 23 and 18 are shown.
  • Additional options for cultivating the characteristics of the piston movement curves of the crank drives provide inclination and displacement of the cylinder axis with respect to their crankshaft axis. Forming the function of the double cranks and rocker arms in addition to a displacement of the cylinder axis of the crank axle (forms eccentric cranks), the movement curve of the piston can be expanded and specified. Shifting the cylinder axis in a double crank causes the connecting rod to get different inclinations to the cylinder axis in TDC and in BDC, so its piston strokes differ from the value of two crank lengths that apply to a centric crank handle.
  • The effect of a displacement and (or) inclination of the cylinder axis relative to the crank axis on the motion function of the piston has been examined here to determine if the displacement or inclination of the cylinder is possible at all when needed and how such action is due to the movement function of the piston effect. The analysis shows that both the displacement and the inclination of the cylinder axis to the crank pin axis bring no disadvantages, but even improve the movement curve of the piston. The effect of the displacement or inclination is greater with mechanisms in the jerk crank mechanism with a smaller distance.
  • In a double crankshaft jerk crank mechanism, when the inclined cylinder axis crosses the axis of the journals of the crank (sole inclination), the oscillations of the connecting rod about its cylinder axis remain symmetrical. Mechanism modified in such a way causes an increased jerk of the piston from the Z-OT to the UT, from where the GW-TDC becomes more fluid. The degree of achievable differences with the original curve with sole inclination without displacement of the cylinder axis depends on the inclination angle. The greater the angle of the cylinder axis to the crankshaft axis the stronger this effect. For example, tilting the cylinder axis to the crankshaft axis 45 ° about the cylinder crank axis for a double crank jerk crank engine 1,25 / 3 causes its 95 percent stroke to be 21 degrees and its 100 percent to be 35 degrees. Without inclination happens in such a construction, a 95-percent stroke in 43 ° and a complete stroke in 180 °. For a double crank jerk crank drive engine 1.5 / 3 (cf. 19, a and c) when the cylinder axis inclines 45 ° about the cylinder crank axis, the 95% downstroke occurs at 38 ° and the full 61 °. Without inclination happens here a 95 percent downstroke in 71 ° and a complete stroke in 180 °. The stroke value is not affected by the cylinder pitch and is always two lengths of the crank (only with rotating cylinder cranks!).
  • The displacement of the axis of the cylinder relative to the crankshaft axis causes similar changes in the piston movement, but this increases the piston stroke in addition and deviates from twice the value of the crank length of the crank handle. With the jerk crank drive engine 1.5 / 3 eg (s. 19, b ), displacement of the cylinder by 1 kl from the crankshaft axis, apart from increasing the stroke, causes the piston movement curve to become more blunt from UT to Z-OT and from Z-OT to UT. Its 95 percent stroke in the power stroke occurs in 42 °, 100 percent downstroke in 76 °, which is not significantly slower than the design with tilt of the cylinder axis by 45 ° (s. 19, a and curves 7 and 8 in FIG 3 ). The course from the Z-OT to the UT hardly changes, the passing speed through the Z-OT also remains almost unchanged. The typical retardation zone of the piston disappears almost completely. After passing through the UT, the piston first moves slowly and relatively evenly about half a turn of the motor shaft, covering about ¼ of its stroke until it gets a smooth acceleration before the last eighth turn of the motor shaft and increases its main ¾ stroke share with increased but again relatively even speed, as in a crank mechanism without displacement of the cylinder axis, reaches the GW-OT.
  • The next example ( 3 ) the design with a centric jerk crank-crank (jerk crank-engine-1.25 / 3, curve 6) shows that realized without displacement of the cylinder axis a 95-percent piston stroke in 43 ° and a full stroke in 180 ° is, in the eccentric variant of the same jerk crank thrust crank (displacement of the cylinder axis by 1 kl of the crank shaft, curve 8), the piston stroke increases by about 7% (2.139 kl), the 95-percent downhill in 24 ° and its full downward stroke from TDC to TDC reaches the piston already at 48 °.
  • Economy of the jerk-crank drive engine.
  • The versatility and flexibility of the jerk-crank mechanism are considerable and it is much better than the conventional crank as a crank mechanism for an internal combustion engine. By varying its wide range selectable distance, the degree of variation in its gear ratio between the coupled cranks during design can be much more accurately tuned to the thermodynamic processes occurring in the engine. If, compared to the HM in a KPM, the original sinusoid of the piston motion generated by the crankshaft is slightly tapered and moderately flattened at UT by the inclination of the connecting rod at TDC, whereby the piston retains its higher velocity longer and throttles it close to UT in the case of the jerk crank drive, the piston curve is sharply sharpened directly at the TDC. The piston gets a jolt here (in the sense of the first derivative of the acceleration after the time). Therefore, the piston acceleration increases rapidly here directly after crossing the TDC. In the jerk crank motor, both the full expansion and its first most important phase with the highest temperature differences of the working gas with the cylinder walls are greatly accelerated. This averts the most significant heat losses during the working cycle, which results in an exceptionally high saving effect. (see diagram in 3 , Curves 6, 7, 8).
  • A jerk crank motor is undoubtedly heavier compared to a simpler design of the same capacity HM, but even then, it still has a much better power to weight ratio. The higher manufacturing costs per product of a jerk crank-type engine are offset by lower manufacturing costs per kW of power and increased economy. Particularly lucrative is a jerk crank engine due to the 12 to 15 times lower NOx emissions and its economy. Investments will soon be fully covered by fuel savings and profits will be made. The jerk-crank-motor construction pays off both by economy and by environmental friendliness highly.
  • reference lists
  • Table list (total 7 sheets with 8 tables).
  • 1. Table 1 Used terms and abbreviations. Second Table 2 Specified parameters of the motors. Third Table 3 Units. 4th Table 4 Degree of fluctuation of the piston speed in the jerk-crank engine. 5th Table 5 Formula table for drawing 20 , 6th Table 6 Unfolding the effective lever arm of the crank for the Pleuelkraft and torque development in the jerk-crank motor in the power stroke. 7th Table 7 Conversion losses, efficiency, power and fuel consumption of the jerk crank engine compared to the HM. 8th. Table 8 Parameters of some jerk crankshaft engines in comparison with the parameters of other engine designs.
  • list of figures
    1. 1. 1 Jerk-crank drive motor (plunger and crosshead version).
    2. Second 1 - 1 Principle of the jerk crank drive based on a four-bar chain.
    3. Third 2 - 1 Jerk-crank drive motor based on a cross-head motor.
    4. 4th 2 - 2 Jerk-crank drive motor based on a plunger engine.
    5. 5th 3 Set of curves of the piston movement of different crank mechanisms.
    6. 6th 4
      1. a. Position φ of the engine crankshaft, at maximum expansion speed of a jerk-crank engine with revolving crank crankshaft as a function of the distance D or the length L of his coupling.
      2. b. Ratio of the top speeds of the expansion of a jerk-crank engine with different distances to that of a HM.
      3. c. Degree of decrease in the heat losses of a jerk-crank drive motor compared to heat losses of an HM in the power stroke as a function of its distance D.
    7. 7th 5 Function of the length of the current effective lever arm of the engine crank for the connecting rod force of the jerk crank mechanism with revolving back-crankshaft as a function of its distance. 6 Function of the instantaneous torque in the power stroke from the position of the engine crankshaft of a jerk crank mechanism with rotating crank crankshaft at different distances.
    8. 8th. 7 4-cylinder Jerk-crank-drive V-engine with offset by 180 ° piston pairs with central engine crankshaft.
    9. 9th 8th 2-cylinder jerk-crank boxer engine with engine crankshaft lying laterally from the crankshaft axis.
    10. 10th 9 Jerk-crank drive motor with gear synchronization of the crankshaft of the cylinder banks.
    11. 11th 10 Device for generating the opposite direction of rotation of the crankshaft to the engine crankshaft from the either-or points (divergent).
    12. 12th 11 Line diagram of the extreme maximum of the transmission ratio in the jerk-crank engine as a function of the distance D between its crankshaft axes of rotation.
    13. 13th 12 Coupling crankshaft gears with centric cranks with engine crankshafts, which lie on the line of the cylinder axis.
    14. 14th 13 Coupling crank gear with laterally offset to the cylinder axis engine crankshaft and coupling crank gear with elementary inclination of the cylinder axis to the crank pin axis.
    15. 15th 14 Construction principle of multi-cylinder engines based on the crank-crank mechanisms.
    16. 16th 15 Back-crank drive with symmetrically swiveled crank arm around its crankshaft axis in the angled from the engine crankshaft sector at an angle 0 <α <360 °.
    17. 17th 16 Back-crank drive with crank-arm swinging symmetrically about its crankshaft axis in the inclined to engine crankshaft sector at angle 0 <α <360 °.
    18. 18th 17 Mechanisms of the crankshaft swinging asymmetrically to the crankshaft axis at an angle 0 <α <180 °.
    19. 19th 18 Diagram of an asymmetric swinging crankshaft crank.
    20. 20th 19 Jerk-crank drive motor with inclined and (or) offset cylinder axis.
    21. 21st 20 Fixed name of the jerk-crank-drive-base-parameter.
    22. 22nd 21 Boxer jerk crank drive engine, cycle cycle.
    23. 23rd 22 Boxer jerk-crank drive engine with staggered cylinder axis, cycle cycle.
    24. 24th 23 Mechanisms of the crank crankshaft oscillating asymmetrically to the crankshaft axis at an angle 0 <α <180 ° with inclined and offset cylinder axes.
    25. 25th 24 Fully balanced 4-cylinder boxer block.
    26. 26th 25 Two-cylinder jerk-crank engine piston engine.
    27. 27th 26 Eight-Cylinder H180 ° - Jerk Cranking Motor with extended radii of the coupling offsets, 135 ° inclined to the cylinder axis Cranking Axis, centered common for two cylinder banks engine crankshaft.
    Figure DE102017003146B3_0027
    Figure DE102017003146B3_0028
    Figure DE102017003146B3_0029
    Figure DE102017003146B3_0030
    Table 5. Formula table for the drawing Fig. 20. Formula symbol, formula Remarks α = π-ω Crank angle of the crankshaft β = ARCTAN (SIN (φ) / (D-COS (φ))) Auxiliary angle 1 γ = 2 × β Coupling tendency to the crankshaft axis δ Inclination angle of the cylinder to Ruckkurbeltriebsachse e Eccentricity of the crank or vertical distance between the axis of the cylinder and the shaft journal axes of the crankshaft. ε compression ratio φ Crank angle of engine crankshaft (argument) ω = π-φ-γ Angle between the engine crank and the coupling λ = R 1/3 of Connecting rod ratio (argument) s = COS (α) + 1- (1-COS (ψ)) / λ Piston stroke from the OT D Distance of the jerk crank mechanism (argument) R = 1 Length of crank of crankshaft (base length) R 1 Length of crankshaft crank (argument) K = D Length of paddock (argument), (exceptions are listed separately) η = 1 - (V1 / V2) 0.4 Theoretical efficiency S = 2πR 2 + (2πR × ((s + 2R / (ε -1)) Area (momentaneous) of the expansion space, kl 2 V = πR 2 × ((s + 2R / (ε -1)) Volume (momentary) of the expansion space, kl 3 A1 = R 1 × SIN (π - α - ψ) Lever arm of the connecting rod force on the crankshaft A2 = R 1 × SIN (φ) Lever arm of the coupling force on the crankshaft A3 = SIN (ω) Lever arm of the coupling force on the engine crankshaft i = R 1 × SIN (φ) / R × SIN (ω) Acute transmission ratio between the crankshafts ψ = ARCSIN (λ × SIN (α)) Inclination angle of the connecting rod to the cylinder axis
    Figure DE102017003146B3_0031
    Table 7. Conversion Losses, Efficiency, Performance, and Fuel Consumption of the Jerk Crank Motor vs. Basic HM. gasoline engine parameter Basic HM Jerk - Crankshaft Engine 1.25 / 3 Limit 1 Limit value 2 Limit 1 Limit value 2 1 degree of expansion 10 10 2 Distance of the crank drive, kl - 1.25 3 Efficiency, theoretically .6019 .6019 4 real statistical efficiency, absolute 0.2500 0.3000 - - 5 complete real losses, absolutely .3519 .3019 - - 6 Dissipation losses, absolutely 0.0500 0.0500 0.0500 0.0500 7 Degree of reduction of heat loss 1.0000 1.0000 3.5606 3.5606 8th Heat loss, absolutely .3019 .2519 0.0848 0.0707 9 Efficiency absolute (calculated) - - .4671 .4811 10 Heat losses, relative 1.0000 1.0000 .2809 .2809 11 Increase in efficiency, relative 1.0000 1.0000 1.8684 1.6038 12 Torque increase, relative 1.0000 1.0000 1.8684 1.6038 13 Max. rel. Speed of the jerk crank engine, in which its cylinder filling of a HM-cylinder filling at max. HM torque remains the same 1.0000 1.0000 1.4089 1.4089 14 Max. Performance increase, relative 1.0000 1.0000 2.6324 2.2596 15 Specific fuel consumption, relative 1.0000 1.0000 .5352 .6235 diesel engine parameter Basic HM Jerk - Crankshaft Engine 1,25 / 3 Limit 1 Limit value 2 Limit 1 Limit value 2 16 degree of expansion 22 22 17 Distance of the crank drive, kl - 1.25 18 Efficiency, theoretically .7096 .7096 19 real statistical efficiency, absolute .3500 .4000 - - 20 complete real losses, absolutely .3596 .3096 - - 21 Dissipation losses, absolutely 0.0500 0.0500 0.0500 0.0500 22 Degree of reduction of heat loss 1.0000 1.0000 3.5606 3.5606 23 Heat loss, absolutely .3096 .2596 .0869 0.0729 24 Efficiency practically absolute (calculated) - - .5726 .5867 25 Heat losses, relative 1.0000 1.0000 .2809 .2809 26 Increase in efficiency, relative 1.0000 1.0000 1.6361 1.4667 27 Torque increase, relative 1.0000 1.0000 1.6361 1.4667 28 Max. rel. Speed of the jerk crank engine, in which its cylinder filling of a HM-cylinder filling at max. HM torque remains the same 1.0000 1.0000 1.4089 1.4089 29 Max. Performance increase, relative 1.0000 1.0000 2.3051 2.0664 30 Specific fuel consumption, relative 1.0000 1.0000 .6112 .6818
    Figure DE102017003146B3_0032
  • Literature:
    1. [1] Author collective under direction of Studienrat Dipl.-Päd. Ing.-Ök. Folkmar Kinzer, knowledge storage engine combustion engines, 6th edition, transpress VEB publishing house for traffic, Berlin 1986 ,
    2. [2] DE102004044214A9 "Short connecting rod crank mechanism" from 14.09.2004.
    3. [3] DE102004057577 "Two-stroke short connecting rod engine" from 30.11.2004.
    4. [4] Works by Prof. Dr.-Ing. Walter Kleinschmidt on transient heat transfer in combustion engines, University of Siegen, http://www.mb.uni-siegen.de/d/ife2/index.htm
    5. [5] Prof. Dr.-Ing. Walter Kleinschmidt, New Theory for Heat Transfer in Internal Combustion Engines, Spectrum of Science, May 1995, pp. 21-30 ,
    6. [6] Efficiency of the gasoline engine, efficiency of the diesel engine, Bibliographisches Institut & F.A. Brockhaus AG, 1999.
    7. [7] DE 86 24 014 U1 "Piston engine".
    8. [8th] DE 603 16 372 T2 "Engine"
    9. [9] FR 900953 "Reciprocating engine with crank mechanism, which interacts with a lever mechanism".
    10. [10] DE 199 42 904 A1 "Internal combustion engine".
  • Jerk-crank drive, as well as equipped with it internal combustion engine, jerk-crank drive engine.
  • Description of Contents:
  • 1.
    introduction
    Second
    Representation of the problem
    Third
    basic idea
    4th
    the solution of the problem
    5th
    Realization of the counterrotation of the crankshafts from the either-or-points
    6th
    Cycle sequence of the jerk crank motor with rotating crank (double crank)
    7th
    Jerk crank mechanism with non-rotatable crankshafts (crank rockers)
    7.1
    Rocker arms with oscillation angles between 0 ° and 180 °
    7.2
    Rocker arms with oscillation angles between 0 and 360 °
    8th.
    Advantages of a jerk crank motor with rotating crankshaft in the cycle compared to the base HM
    9th
    Valve timing
    10th
    The calculation of the increase in the efficiency of the jerk crank engine
    11th
    Calculation of the power increase of the jerk crank engine
    12th
    Environmental compatibility and emissions
    13th
    Embodiments of the jerk crank drive motor
    14th
    Opposite piston jerk crank drive motor
    15th
    Additional optimization options for the piston movement of a jerk-crank drive motor
    16th
    Economy of the jerk-crank drive engine
    17th
    reference lists
    17.1
    table list
    17.2
    List of drawings and diagrams
    18th
    literature

    Claims (12)

    1. Jerk crank mechanism, consisting of the engine crankshaft (5), which with its shaft journals (5/1) in the bearings (14) of the motor housing (13), the return crankshaft (6), with its shaft journals (6/1) in the bearings (15) of the motor housing (13), the coupling (7) and the piston (2), which in the cylinder (1), which is fixedly installed in the motor housing (13) is axially displaceable, and either by means of Bolzens (4) directly to the connecting rod eye of the connecting rod (3), is rotatably coupled, or indirectly, by means of the fixedly connected rod (10) and, in the permanently installed in the motor housing (13) Gleitlagerwandungen (12) via its sliding shoe ( 11) rectilinearly displaceable crosshead (8) which is rotatably coupled by means of the bolt (4) on the connecting rod eye of the connecting rod (3), characterized in that the connecting rod (3) at its connecting rod by means of the coupling (7) on the crank pin (5 / 2) of the engine crankshaft (5) is rotatably coupled and the connecting rod coupling joint (3-7) of the crank pin (6/2) d he crankshaft shaft (6) is guided around its shaft journals (6/1);
    2. Jerk-crank mechanism after Claim 1 , characterized in that its cylinder axis to its crankshafts an arbitrary position which coincides with the normal distance e to the axis of rotation of the crankshaft (6) and the inclination angle δ to the plane in which the axis of rotation of the shaft journals (6/1) of the return crankshaft (6 ) and the axis of rotation of the shaft journals (5/1) of the engine crankshaft (5) lie, describes, has;
    3. Jerk crank mechanism according to one of the preceding claims, characterized in that the plane formed by the axis of the shaft journals (6/1) of its return crankshaft (6) and of the axis of the shaft journals (5/1) of its engine crankshaft (5) is at an angle δ to the cylinder axis and the Kröpfungsebene of the crank pin (6/2) of the return crankshaft (6) to which the connecting rod of the connecting rod (3) is coupled to the cranking plane of its crank pin (6/3), to which the return crankshaft (6) is coupled by means of the coupling (7) to the crank pin (5/2) of the engine crankshaft (5), offset by the same angle δ about the axis of its shaft journals (6/1);
    4. Jaw crankshaft drive according to one of the preceding claims, characterized in that the radius R1 of the cranking of the return crankshaft (6) is equal to the radius R of the cranking of the engine crankshaft (5) (R1 = R), the length K of its coupling (7) Distance D between the axis of the shaft journal (5/1) of the engine crankshaft (5) and the axis of the shaft journal (6/1) of the return crankshaft (6) is equal to (K = D) and the connecting rod coupling joint (3-7 ) is guided by the crank pin (6/2) of the return crankshaft (6) about the axis of its shaft journals (6/1) against the direction of rotation of the engine crankshaft (5) in circles;
    5. Jerk-crank mechanism after one of the Claims 1 to 3 , characterized in that the radius R1 of the cranking of the return crankshaft (6) is greater than the radius R of the crank of the engine crankshaft (5) (R1> R), the length K of the coupling (7) by the difference of the crankshaft crankshaft (R1 - R). R) is smaller than the distance D between the axis of the shaft journals (5/1) of the engine crankshaft (5) and the axis of the shaft journals (6/1) of the back crankshaft (6) (K = D - (R1 - R)) and Connecting rod coupling joint (3-7) in the arc around the axis of the shaft journal (6/1) of the return crankshaft (6) at an angle of 0 ° <α <360 ° oscillates;
    6. Jerk-crank mechanism after one of the Claims 1 to 3 , characterized in that the radius R1 of the cranking of the return crankshaft (6) is greater than the radius R of the crank of the engine crankshaft (5) (R1> R), the length K of the coupling (7) by the difference of the crankshaft crankshaft (R1 - R). R) is greater than the distance D between the shaft journal shaft (5/1) of the engine crankshaft (5) and the shaft journal shaft (6/1) of the crankshaft (6) (K = D + (R1 - R)) and Connecting rod coupling joint (3-7) in the arc around the axis of the shaft journal (6/1) of the return crankshaft (6) at an angle of 0 ° <α <360 ° oscillates;
    7. Jerk-crank mechanism after one of the Claims 1 to 3 , characterized in that the radius R1 of the crank of the crankshaft (6) is greater than the radius R of the crank of the engine crankshaft (5) (R1> R), the length K of the coupling (7) of the distance D between the axis of the shaft journals (5/1) of the engine crankshaft (5) and the axis of the shaft journals (6/1) of the return crankshaft (6) is equal to (K = D) and the connecting rod coupling joint (3-7) in the arc around the axis of Shaft journal (6/1) of the return crankshaft (6) oscillates at an angle of 0 ° <α <180 °;
    8. Jerk-crank mechanism after one of the Claims 4 to 7 CHARACTERIZED IN THAT, if its predetermined distance D can not be achieved in a real jerk crank drive design to maintain the intended jerk crank drive function parameters, the real parameters of the following limbs of its kinematic chain will take such values: the length of the coupling (7): K kopp. = K * K, the radius of the coupling cranking of the engine crankshaft (5): R kopp. = K · R the radius of the coupling cranking of the crankshaft (6): R1 kopp. = k · R1, where k is the ratio of the real distance D real to the given D: k = D real : D;
    9. Jerk crank mechanism according to one of the preceding claims, characterized in that for generating the opposite starting direction of rotation of its return crankshaft (s) to the (n) engine crankshaft (s) when starting from the either-or-points, he with device (s) ( Divergator (s)) for being equipped;
    10. Jerk-crank mechanism after the Claim 9 , characterized in that its divergator represents a single-stage gear transmission with gears with plausible module in the form of elliptical sectors - one with at least one tooth, the other with at least one tooth gap - whose cyclically fluctuating gear ratio is congruent with that of the jerk crank mechanism, it reaches extreme gear ratio at the either-or-point and covers a sufficient symmetrical to either-or-point tooth contact area;
    11. Combustion engine according to two-, four-stroke or split-cycle process, with external or internal mixture formation, with self-ignition or spark ignition, suction or supercharger motor of any designs, consisting of at least one cylinder, characterized in that it with at least one jerk crank mechanism with at least a crank handle is equipped according to at least one of the preceding claims;
    12. Internal combustion engine Claim 11 , characterized in that its valve and ignition timing to the cycle sequences of its crank mechanism construction with their specific parameters, are logically adjusted.
    DE102017003146.0A 2017-03-30 2017-03-30 Jerk-crank mechanism, as well as equipped combustion engine. Active DE102017003146B3 (en)

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    Cited By (2)

    * Cited by examiner, † Cited by third party
    Publication number Priority date Publication date Assignee Title
    RU2716521C1 (en) * 2019-07-30 2020-03-12 Общество с ограниченной ответственностью "Завод дозировочной техники "Ареопаг" Piston device of pump
    DE102019004694B3 (en) * 2019-07-03 2020-09-10 Georg Schreiber Planetary gear for pendulum journals

    Citations (6)

    * Cited by examiner, † Cited by third party
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