Construction of indicator charts
Indicator diagrams are built in coordinates p-V.
The construction of an indicator diagram of an internal combustion engine is based on a thermal calculation.
At the beginning of construction, on the abscissa axis, a segment AB is plotted, corresponding to the working volume of the cylinder, and in magnitude equal to the piston stroke on a scale, which, depending on the piston stroke of the designed engine, can be taken as 1:1, 1.5:1 or 2:1.
Segment OA, corresponding to the volume of the combustion chamber,
is determined from the ratio:
Segment z "z for diesel engines (Fig. 3.4) is determined by the equation
Z,Z=OA(p-1)=8(1.66-1)=5.28mm, (3.11)
pressures = 0.02; 0.025; 0.04; 0.05; 0.07; 0.10 MPa in mm so that
get the height of the chart equal to 1.2 ... 1.7 of its base.
Then, according to the thermal calculation data on the diagram, they are laid in
the chosen scale of pressure values at the characteristic points a, c, z", z,
b, r. The z point for a gasoline engine corresponds to pzT.
Four-stroke diesel engine indicator diagram
According to the most common Brouwer graphical method, compression and expansion polytropes are built as follows.
Draw a ray from the origin OK at an arbitrary angle to the abscissa axis (it is recommended to take = 15 ... 20 °). Further, from the origin, rays OD and OE are drawn at certain angles and to the y-axis. These angles are determined from the relations
0.46 = 25°, (3.13)
The compression polytrope is built using the rays OK and OD. From point C, a horizontal line is drawn until it intersects with the y-axis; from the intersection point - a line at an angle of 45 ° to the vertical until it intersects with the OD beam, and from this point - a second horizontal line parallel to the abscissa axis.
Then a vertical line is drawn from point C until it intersects with the OK beam. From this point of intersection at an angle of 45 ° to the vertical, we draw a line until it intersects with the abscissa axis, and from this point?? a second vertical line parallel to the y-axis, until it intersects with the second horizontal line. The intersection point of these lines will be the intermediate point 1 of the compression polytrope. Point 2 is found similarly, taking point 1 as the beginning of the construction.
The expansion polytrope is built using the rays OK and OE, starting from the point Z", similar to the construction of the compression polytrope.
The criterion for the correct construction of the extension polytrope is its arrival at the previously plotted point b.
It should be borne in mind that the construction of the expansion polytropic curve should be started from the point z , and not z ..
After constructing the contraction and expansion polytropes, they produce
rounding the indicator diagram taking into account the pre-opening of the exhaust valve, the ignition timing and the rate of pressure rise, and also apply the intake and exhaust lines. For this purpose, under the abscissa axis, a semicircle with radius R=S/2 is drawn on the piston stroke length S as on the diameter. From the geometric center Оґ in the direction of n.m.t. a segment is postponed
Where L- the length of the connecting rod, is selected from the table. 7 or prototype.
Ray ABOUT 1.WITH 1 is carried out at an angle Q o = 30° corresponding to the angle
ignition timing ( Qo= 20…30° to w.m.t.), and the point WITH 1 demolished on
contraction polytrope, obtaining the point c1.
To build lines for cleaning and filling the cylinder, a beam is laid ABOUT 1?IN 1 at an angle g=66°. This angle corresponds to the pre-opening angle of the exhaust valve or exhaust ports. Then a vertical line is drawn until it intersects with the expansion polytrope (point b 1?).
From a point b 1. draw a line that defines the law of change
pressure in the section of the indicator diagram (line b 1.s). Line as,
characterizing the continuation of cleaning and filling the cylinder, can
be held straight. It should be noted that the points s. b 1. you can also
find by the value of the lost fraction of the piston stroke y.
as=y.S. (3.16)
The indicator diagram of two-stroke engines, as well as supercharged engines, always lies above the atmospheric pressure line.
In a supercharged engine indicator chart, the intake line may be higher than the exhaust line.
The indicator diagram of the internal combustion engine (Fig. 1) is built using the calculation data of the processes of the engine working cycle. When constructing a diagram, it is necessary to choose a scale in such a way as to obtain a height equal to 1.2 ... 1.7 of its base.
Fig.1 Diesel engine indicator diagram
Rice. 1 Diesel engine indicator diagram
At the beginning of the construction, on the abscissa axis (the base of the diagram), the segment S a \u003d S c + S is plotted on the scale,
where S is the stroke of the piston (from TDC to BDC).
Segment S c corresponding to the volume of the compression chamber (V c) is determined by the expression S c = S / - 1.
The segment S corresponds to the working volume V h of the cylinder, and is equal in magnitude to the piston stroke. Mark the points corresponding to the position of the piston at TDC, points A, B, BDC.
The pressure on the scale of 0.1 MPa per millimeter is plotted along the ordinate axis (diagram height).
Pressure points p g, p c, p z are plotted on the TDC line.
Pressure points p a, p c are plotted on the NDC line.
For a diesel engine, it is also necessary to plot the coordinates of the point corresponding to the end of the calculated combustion process. The ordinate of this point will be equal to p z, and the abscissa is determined by the expression
S z = S with , mm. (2.28)
The construction of the line of compression and expansion of gases can be carried out in the following sequence. Arbitrarily, between TDC and BDC, at least 3 volumes or segments of the piston stroke V x1, V x2, V x3 (or S x1, S x2, S x3) are selected.
And gas pressure is calculated
On the compression line
On the expansion line
All constructed points are smoothly connected to each other.
Then the transitions are rounded off (with each change in pressure at the junctions of the calculated cycles), which is taken into account in the calculations by the coefficient of completeness of the diagram.
For carburetor engines, the rounding at the end of combustion (point Z) is carried out along the ordinate p z \u003d 0.85 P z max.
2.7 Determining the mean indicator pressure from the indicator chart
The average theoretical indicator pressure p "i is the height of a rectangle equal to the area of the indicator diagram in the pressure scale
MPa (2.31)
where F i is the area of the theoretical indicator diagram, mm 2, limited by the lines of TDC, BDC, compression and expansion, can be determined using a planimeter, by the integration method, or in another way; S - indicator diagram length (piston stroke), mm (distance between TDC, BDC lines);
p - pressure scale selected when constructing the indicator diagram, MPa / mm.
Actual indicator pressure
р i = р i ΄ ∙ φ p, MPa, (2.32)
where p - coefficient of incompleteness of the area of the indicator diagram; takes into account the deviation of the actual process from the theoretical one (rounding with a sharp change in pressure, for carburetor engines p = 0.94 .. .0.97; for diesel engines p = 0.92 .. .0.95);
р = р r - ра - average pressure of pumping losses during intake and exhaust for naturally aspirated engines.
After determining p i according to the indicator diagram, it is compared with the previously calculated one (formula 1.4) and the discrepancy is determined as a percentage.
Mean effective pressure p e is equal to
p e \u003d p i - p mp,
where p mp is determined by formula 1.6.
Then calculate the power according to the dependence 
and compare with the given one. The discrepancy should be no more than 10 ... 15%, if more processes should be recalculated.
It is advisable to study the operation of a real piston engine using a diagram that shows the change in pressure in the cylinder depending on the position of the piston for the entire
cycle. Such a diagram, taken using a special indicator device, is called an indicator diagram. The area of the closed figure of the indicator diagram depicts on a certain scale the indicator work of the gas in one cycle.
On fig. Figure 7.6.1 shows the indicator diagram of an engine operating with fast-burning fuel at constant volume. As a fuel for these engines, light fuel gasoline, lighting or generator gas, alcohols, etc. are used.
When the piston moves from the left dead position to the extreme right, a combustible mixture is sucked in through the suction valve, consisting of vapors and small particles of fuel and air. This process is depicted in a 0-1 curve diagram, which is called the suction line. Obviously, the 0-1 line is not a thermodynamic process, since the main parameters do not change in it, but only the mass and volume of the mixture in the cylinder change. When the piston moves back, the suction valve closes, and the combustible mixture is compressed. The compression process in the diagram is depicted by a curve 1-2, which is called the compression line. At point 2, when the piston has not yet reached the left dead position, the combustible mixture is ignited by an electric spark. Combustion of the combustible mixture occurs almost instantly, i.e., almost at a constant volume. This process is depicted in the diagram by curve 2-3. As a result of fuel combustion, the gas temperature rises sharply and the pressure increases (point 3). Then the combustion products expand. The piston moves to the right dead position, and the gases do useful work. On the indicator diagram, the expansion process is depicted by a 3-4 curve, called the expansion line. At point 4, the exhaust valve opens and the pressure in the cylinder drops to almost outside pressure. With further movement of the piston from right to left, combustion products are removed from the cylinder through the exhaust valve at a pressure slightly higher than atmospheric pressure. This process is depicted in the 4-0 curve diagram and is called the exhaust line.
The considered working process is completed in four strokes of the piston (cycle) or in two revolutions of the shaft. Such engines are called four-stroke.
From the description of the operation of the process of a real internal combustion engine with rapid combustion of fuel at a constant volume, it can be seen that it is not closed. It has all the signs of irreversible processes: friction, chemical reactions in the working fluid, final piston speeds, heat transfer at a finite temperature difference, etc.
Consider an ideal thermodynamic cycle of an engine with an isochoric supply of heat (v=const), consisting of two isochores and two adiabats.
On fig. 70.2 and 70.3 show a cycle in - and - diagrams, which is carried out as follows.
An ideal gas with initial parameters and is compressed along the adiabatic 1-2 to point 2. The amount of heat is reported to the working fluid along the isochore 2-3. From point 3, the working body expands along the adiabatic 3-4. Finally, along the 4-1 isochore, the working fluid returns to its original state, while the amount of heat is removed to the heat sink. The characteristics of the cycle are the compression ratio and the pressure ratio.
We determine the thermal efficiency of this cycle, assuming that the heat capacity and value are constant:
The amount of heat supplied and the amount of heat removed.
Then the thermal efficiency of the cycle
Rice. 7.6.2 Fig. 7.6.3
Thermal efficiency of a cycle with heat input at constant volume
. (7.6.1) (17:1)
From equation (70.1) it follows that the thermal efficiency of such a cycle depends on the degree of compression and the adiabatic index or on the nature of the working fluid. The efficiency increases with increasing and . From the degree of pressure increase, thermal efficiency does not depend.
Taking into account - diagrams (Fig. 70.3), the efficiency is determined from the ratio of areas:
\u003d (pl. 6235-pl. 6145) / square. 6235 = pl. 1234/pl. 6235.
It is very clearly possible to illustrate the dependence of the efficiency on the increase in the - diagram (Fig. 7.70.3).
If the areas of the supplied amount of heat are equal in two cycles (pl. 67810 = pl. 6235), but at different degrees of compression, the efficiency will be greater for the cycle with a higher degree of compression, since a smaller amount of heat is removed to the heat sink, i.e. pl. 61910<пл. 6145.
However, an increase in the compression ratio is limited by the possibility of premature self-ignition of the combustible mixture, which disrupts the normal operation of the engine. In addition, at high compression ratios, the rate of combustion of the mixture increases dramatically, which can cause detonation (explosive combustion), which dramatically reduces the efficiency of the engine and can lead to breakage of its parts. Therefore, a certain optimal compression ratio must be applied for each fuel. Depending on the type of fuel, the compression ratio in the studied engines varies from 4 to 9.
Thus, studies show that high compression ratios cannot be used in internal combustion engines with constant volume heat input. In this regard, the considered engines have relatively low efficiency.
The theoretical useful specific work of the working fluid depends on the relative position of the processes of expansion and contraction of the working fluid. Increasing the average pressure difference between the expansion and compression lines makes it possible to reduce the size of the engine cylinder. If we denote the average pressure through then the theoretical useful specific work of the working fluid will be
The pressure is called the average indicator pressure (or average cycle pressure), that is, it is a conditional constant pressure, under the influence of which the piston performs work during one stroke equal to the work of the entire theoretical cycle.
Cycle with the supply of the amount of heat in the process
The study of cycles with the supply of heat at a constant volume showed that in order to increase the efficiency of an engine operating according to this cycle, it is necessary to use high compression ratios. But this increase is limited by the self-ignition temperature of the combustible mixture. If, however, to produce separate compression of air and fuel, then this restriction disappears. Air at high compression has such a high temperature that the fuel supplied to the cylinder spontaneously ignites without any special ignition devices. And finally, separate compression of air and fuel allows the use of any liquid heavy and cheap fuel - oil, fuel oil, resins, coal oils, etc.
Such high advantages are possessed by engines operating with gradual combustion of fuel at constant pressure. In them, air is compressed in the engine cylinder, and liquid fuel is sprayed with compressed air from the compressor. Separate compression allows the use of high compression ratios (up to ) and eliminates premature self-ignition of the fuel. The process of burning fuel at a constant pressure is ensured by the appropriate adjustment of the fuel injector. The creation of such an engine is associated with the name of the German engineer Diesel, who first developed the design of such an engine.
Consider an ideal engine cycle with gradual combustion of fuel at constant pressure, i.e., a cycle with the supply of heat at constant pressure. On fig. 70.4 and 70.5 this cycle is shown in diagrams. It is carried out as follows. The gaseous working fluid with initial parameters , , is compressed along the adiabatic 1-2; then a certain amount of heat is imparted to the body along the 2-3 isobar. From point 3, the working body expands along the adiabatic 3-4. And finally, along the 4-1 isochore, the working fluid returns to its original state, while heat is removed to the heat sink.
The characteristics of the cycle are the compression ratio and the pre-expansion ratio.
Let us determine the thermal efficiency of the cycle, assuming that the heat capacities and and their ratio are constant:
The amount of heat supplied
amount of heat removed
Thermal cycle efficiency
Rice. 7.6.4 Fig. 7.6.5
The average indicator pressure in the cycle with heat supply at is determined from the formula
The mean indicator pressure increases with increasing and .
A cycle with the supply of heat in the process at and , or a cycle with a mixed supply of heat.
Engines with gradual combustion of fuel at have some disadvantages. One of them is the presence of a compressor used to supply fuel, the operation of which consumes 6–10% of the total engine power, which complicates the design and reduces the efficiency of the engine. In addition, it is necessary to have complex pump devices, nozzles, etc.
The desire to simplify and improve the operation of such engines has led to the creation of compressorless engines in which fuel is mechanically atomized at pressures of 50–70 MPa. The project of a compressorless high-compression engine with a mixed supply of heat was developed by the Russian engineer G.V. Trinkler. This engine is devoid of the shortcomings of both disassembled engine types. Liquid fuel is supplied by the fuel pump through the fuel injector to the cylinder head in the form of tiny droplets. Getting into the heated air, the fuel spontaneously ignites and burns during the entire period while the nozzle is open: first at a constant volume, and then at a constant pressure.
The ideal cycle of an engine with a mixed heat input is shown in - and - diagrams in fig. 70.6 and 70.7.
.Let us determine the thermal efficiency of the cycle, provided that the heat capacities , and the adiabatic exponent are constant:
The first fraction of the amount of heat supplied
The second share of the amount of heat supplied
The amount of heat removed
The indicator diagram - the dependence of the pressure of the working fluid on the volume of the cylinder (Fig. 2) - is the most informative source that allows you to analyze the processes occurring in the cylinder of an internal combustion engine. The engine cycles, carried out in four piston strokes from TDC to BDC, are shown on the indicator diagram in the coordinates p–V the following curve segments:
r 0 – a 0 - intake stroke;
a 0 – c- compression stroke;
c – z-b 0 – cycle of the working stroke (expansion);
b 0 – r 0 – release stroke.
The following characteristic points are marked on the diagram:
b, r- the opening and closing times of the exhaust valve, respectively;
u, a - the opening and closing times of the intake valve, respectively;
Rice. 2. Typical indicator diagram of a four-stroke
internal combustion engine
The area of the diagram that determines the work per cycle consists of the area corresponding to the positive indicator work obtained during the compression and stroke strokes, and the area corresponding to the negative work spent on cleaning and filling the cylinder in the intake and exhaust strokes. Negative cycle work is usually referred to as mechanical losses in the engine.
Thus, the total energy imparted to the piston engine shaft in one cycle L, can be determined by algebraic addition of the work of cycles L = L ch + L szh + L px + L issue The power transmitted to the shaft will be determined by the product of this sum by the number of cycles of the working stroke per unit of time ( n/2) and on the number of engine cylinders i:
The engine power determined in this way is called the average indicated power.
The indicator diagram allows you to divide the cycle of a four-stroke engine into the following processes:
u –r 0 – r – a 0 -a- inlet;
a – θ – c" – compression;
θ – c" – c – z – f – mixture formation and combustion;
z-f-b- extension;
b –b 0 – u – r 0 – r – release.
The typical indicator diagram shown is also valid for a diesel engine. In this case, the point θ will correspond to the moment of fuel supply to the cylinder.
The diagram shows:
V c – combustion chamber volume (cylinder volume above the piston at TDC);
Va- gross volume of the cylinder (the volume of the cylinder above the piston at the beginning of the compression stroke);
V n – working volume of the cylinder, V n = V a – V c.
Compression ratio.
The indicator diagram describes the operating cycle of the engine, and its limited area – cycle indicator work. Really, [ p ∙ ∆V] \u003d (N / m 2) ∙ m 3 \u003d N ∙ m \u003d J.
If we assume that a certain conditional constant pressure acts on the piston p i , performing during one stroke of the piston work equal to the work of gases per cycle L, That
L = p i ∙ V h()
Where V h is the working volume of the cylinder.
This conditional pressure p i called the mean indicator pressure.
The average indicator pressure is numerically equal to the height of a rectangle with a base equal to the working volume of the cylinder V h with an area equal to the area corresponding to the work L.
Since the useful indicator work is proportional to the average indicator pressure p i , the perfection of the working process in the engine can be evaluated by the value of this pressure. The more pressure p i , the more work L, and hence the working volume of the cylinder is better utilized.
Knowing the average indicator pressure p i , working volume of the cylinder V h , number of cylinders i and crankshaft speed n(rpm), you can determine the average indicated power of a four-stroke engine N i
Work i ∙ V h is the displacement of the engine.
The transfer of the indicator power to the engine shaft is accompanied by mechanical losses due to friction of the pistons and piston rings against the cylinder walls, friction in the bearings of the crank mechanism. In addition, part of the indicator power is spent on overcoming aerodynamic losses that occur during the rotation and oscillation of parts, on actuating the gas distribution mechanism, fuel, oil and water pumps and other auxiliary engine mechanisms. Part of the indicator power is spent on removing combustion products and filling the cylinder with a fresh charge. The power corresponding to all these losses is called the power of mechanical losses. N m.
In contrast to the indicated power, the useful power that can be obtained on the motor shaft is called effective power. N e. The effective power is less than the indicator power by the amount of mechanical losses, i.e.
N e = N i- N m. ()
Power N m corresponding to mechanical losses and effective engine power N e is determined empirically during bench tests using special load devices.
One of the main indicators of the quality of a piston engine, which characterizes the use of indicator power by it to perform useful work, is mechanical efficiency, defined as the ratio of effective power to indicator power:
η m = N e / N i . ()
The total energy imparted to the shaft of a piston engine can be determined by algebraic addition of the work cycles and multiplying the sum by the number of work cycles per unit time ( n/2) and the number of engine cylinders. The power determined in this way can be obtained by integrating the dependence of pressure as a function of volume shown in the indicator diagram (Figure 4.2, b), and is called the average indicator power N. This power is often associated with the concept of indicator mean effective pressure R i , calculated as follows:
Effective power N e is the product of the indicator power N on the mechanical efficiency of the engine. The mechanical efficiency of the engine decreases with increasing engine speed due to friction losses and drive units.
To build the characteristics of an aircraft piston engine, it is tested on a balancing machine using a variable pitch propeller. The balancing machine provides measurement of the torque, the number of revolutions of the crankshaft and fuel consumption. According to the measured torque M kr and number of revolutions n the measured effective motor power is determined
If the engine is equipped with a gearbox that reduces the speed of the propeller, then the formula for the measured effective power is:
Where i p is the gear ratio of the gearbox.
Taking into account the dependence of the effective engine power on atmospheric conditions, the measured power for comparison of test results is reduced to standard atmospheric conditions according to the formula
Where N e is the effective engine power reduced to standard atmospheric conditions;
t meas - outdoor air temperature during testing, ºС;
B- outside air pressure, mm Hg,
∆R– absolute air humidity, mm Hg.
Effective specific fuel consumption g e is determined by the formula:
Where G T and - fuel consumption and effective engine power, measured during tests.
The main difference between a 2-stroke engine and a 4-stroke one is the method of gas exchange - cleaning the cylinder from combustion products and charging it with fresh air or a hot mixture.
Gas distribution devices of 2-stroke engines - slots in the cylinder liner, blocked by a piston, and valves or spools.
Duty Cycle:
After the combustion of the fuel, the process of expansion of gases (working stroke) begins. The piston moves to bottom dead center (BDC). At the end of the expansion process, piston 1 opens inlet slots (windows) 3 (point b) or exhaust valves open, communicating the cylinder cavity through the exhaust pipe with the atmosphere. In this case, part of the combustion products leaves the cylinder and the pressure in it drops to the purge air pressure Pd. At point d, the piston opens purge windows 2, through which a mixture of fuel and air is supplied to the cylinder at a pressure of 1.23-1.42 bar. Further fall slows down, because. air enters the cylinder. From point d to BDC, outlet and purge windows are simultaneously open. The period during which the purge and exhaust ports are open at the same time is called purge. During this period, the cylinder is filled with a mixture of air, and the combustion products are displaced from it.
The second stroke corresponds to the piston stroke from bottom to top dead center. At the beginning of the stroke, the purge process continues. Point f - the end of the purge - the closing of the inlet windows. At point a, the outlet windows close and the compression process begins. The pressure in the cylinder at the end of charging is slightly higher than atmospheric pressure. It depends on the purge air pressure. From the moment the purge is completed and the exhaust windows are completely closed, the compression process begins. When the piston does not reach 10-30 degrees along the angle of rotation of the crankshaft to TDC (point c /), fuel is supplied to the cylinder through the nozzle or the mixture is ignited and the cycle repeats.
With the same cylinder dimensions and rotational speed, the power of the 2-stroke is much greater, 1.5-1.7 times.
Average pressure of the theoretical ICE diagram.
The average indicator pressure of the internal combustion engine.
This is such a conditionally constant pressure, which, acting on the piston, does work equal to the internal work of the gas throughout the entire working cycle.
Graphically, p i on a certain scale is equal to the height of the rectangle mm / hh / , equal in area to the area of the diagram and having the same length.
f- area of the indicator diagram (mm 2)
l- length of index diagram - mh
k p - pressure scale (Pa/mm)
Average effective pressure of internal combustion engine.
This is the product of the mechanical efficiency and the average indicator pressure.
Where η mech =N e /N i . During normal operation η mech =0.7-0.85.
Mechanical efficiency of the internal combustion engine.
η fur \u003d N e / N i
Ratio of effective power to indicator power.
During normal operation η mech =0.7-0.85.
The indicator power of the internal combustion engine.
Ind. the engine power received inside the wheeled wheel can be determined using an indicator diagram taken by a special device - an indicator.
Ind.power - the work done by the working fluid in the engine cylinder in a unit of time.
Individual power of one cylinder -
k- engine power
V-cylinder displacement
n is the number of working moves.
The effective power of the internal combustion engine.
Useful power taken from the crankshaft
N e \u003d N i -N tr
N tr - the sum of power losses due to friction between moving parts of the engine and to actuate auxiliary mechanisms (pumps, generator, fan, etc.)
The determination of the effective power of the engine in laboratory conditions or during bench tests is carried out using special braking devices - mechanical, hydraulic or electrical.
