| WO2001049974A1 | 2001-07-12 | |||
| WO2000023699A1 | 2000-04-27 |
| DE20214880U1 | 2003-01-02 | |||
| US5076220A | 1991-12-31 | |||
| JPS60256642A | 1985-12-18 | |||
| FR483083A | 1917-05-25 | |||
| US4945866A | 1990-08-07 | |||
| DE4318401A1 | 1994-12-08 | |||
| US6612281B1 | 2003-09-02 | |||
| JP2003035146A | 2003-02-07 | |||
| DE10222905A1 | 2003-12-24 | |||
| JP2001182550A | 2001-07-06 | |||
| DE10052350A1 | 2002-05-02 | |||
| CN1363766A | 2002-08-14 |
CRITERIA OF OPTIMIZATION
PATENT CLAIMS
1) The OPTIMIZATION OF HEAT-ENGINES, as defined above (Title, Summary, Description, Drawings), of which we CLAIM all APPLICATIONS and MODIFICATIONS, concerns an ABSOLUTE NOVELTY actually INCOMPATIBLE with the SECOND-LA W-OF-THERMODYNAMICS.
2) The OPTIMIZATION OF HEAT-ENGINES as in the previous CLAIM, characterized by the fact that while OPERATING (after the engine has been started), every Internal-Combustion-HE AT-ENGINE FUNCTIONS AUTONOMOUSLY, INDEPENDENTLY, as a "CLOSED-BOX", thus SELF-REGULATING in REAL-TIME its ENERGETIC-BALANCE, that is the THERMAL EXCHANGE (δQ) and the MECHANICAL EXCHANGE (δL) with the exterior, also SELF-REGULATING its EFFICIENCY.
3) The OPTIMIZATION OF HEAT-ENGINES as in the previous CLAIMS, characterized by the fact that ALTERNATIVE-HEAT-ENGINES are not more economical than TURBINE-ENGINES of equal POWER, even though their IDEAL-CYCLE EFFICIENCIES are higher, while their FRICTIONS are almost IDENTICAL. The DISCREPANCY depends on the MECHANICAL DIFFERENCE between the two ENGINES, given that in TURBINE ENGINES the FORCES act on the DRIVE SHAFT while in ALTERNATIVE-HEAT-ENGINES they become concentrated at DEAD-POINTS, where they are not useful.
4) The OPTIMIZATION OF HEAT-ENGINES as in the previous CLAIMS, characterized by the fact that the EFFICIENCY of the Internal-Combustion-HEAT-ENGINES may be increased significantly by OPTIMIZING their POWER-TRAIN-ASSEMBLIES, MODIFYING opportunely the MECHANISMS that transmit the THERMAL- FORCES of the GAS in COMBUSTION CHAMBER to the DRIVE SHAFT.
5) The OPTIMIZATION OF HEAT-ENGINES as in the previous CLAIMS, characterized by the fact that in TURBINE ENGINES an improvement in EFFICIENCY may be obtained by MODIFYING opportunely the VANES of the ROTORS until the LOCOMOTIVE-FORCES act as far as possible at a TANGENT to the DRIVE SHAFT.
6) The OPTIMIZATION OF HEAT-ENGINES as in the previous CLAIMS, characterized by the fact that all ALTERNATIVE-ENGINES are actioned by the CRANK-CONNECTING-ROD-MECHANISM, where Thermal Forces are concentrated at DEAD-POINTS, where they are not useful. Therefore the OPTIMIZATION of the respective EFFICIENCIES, which may be increased by as much as 30% and over, is of great interest. CRITERIA OF OPTIMIZATION
7) The OPTIMIZATION OF HEAT-ENGINES as in the previous CLAIMS, characterized by the fact that the OPTIMIZATION of ALTERNATIVE-ENGINES is obtained (fig. 6) by displacing the DRIVE SHAFT by a maximum distance h=|OH| from the AXIS of the CYLINDER, after having SET the positive sense of ROTATION according to the new PHASE of EXPANSION (%+aι)≤φ≤(2π+a 2 ), greater than the successive COMPRESSION α 2 <φ≤(π+αi).
8) The OPTIMIZATION OF HEAT-ENGINES as in the previous CLAIMS, characterized by the fact (fig. 6) that in order to assure the CONTINUITY of the MOVEMENT Vφ>0 we should assign the EXTENSION λ=b/r and the ECCENTRICITY p=h/r in the Intervals (λ≥2) and [p<(λ-l)], that exclude the SINGULARITIES (λ<2) and (p≥l).
9) The OPTIMIZATION OF HEAT-ENGINES as in the previous CLAIMS, characterized by the fact (fig. 7) that, excluding the SINGULARITIES (λ<2) and (p≥l), the MAXIMUM-EFFICIENCY is obtained (10) by assigning the MAXIMUM-EXPANSION (α,+π)<φ<(2π+α 2
) and the MINIMUM-COMPRESSION α 2
<φ<(π+cti), so that
10) The OPTIMIZATION OF HEAT-ENGINES as in the previous CLAIMS, characterized by the fact that in ECCENTRIC-ENGINES (h=|OH|>0) VIBRATIONS can be REDUCED by coupling two or more COAXIAL- CYLINDERS in PARALLEL with opposing PHASES, apart from possible FLY-WHEELS. FRICTIONS, produced in the CYLINDER due to the ECCENTRICITY p<(λ-l) and Vλ>2, can also be REDUCED by connecting the Piston-Rod to the MECHANISM through a CROSS-SLIDE-HEAD as in some STEAM ENGINES. |
CRITERIA OF OPTIMIZATION
TITLE
DESCRIPTION of the Industrial Invention entitled:
"CRITERIA OF OPTIMIZATION IN THE EFFICIENCY OF A FIRST HEAT ENGINE",
DESCRIPTION
The REAL-EFFICIENCY η R
=L R
/Q R
of any INTERNAL-COMBUSTION-HEAT-ENGINE depends on the IDEAL- CYCLE 3 (Otto, Diesel, Joule, Carnot, etc.) CREATED (in whichever way) in the Combustion Chamber, in addition to the INTERNAL-EFFICIENCY η o
=(l-λ A
)(l-λc) determined by FRICTIONS (λ A
λc), the MECHANICAL EFFICIENCY η A
=(l-λ A
)=L R
/L (which transforms Work into Heat by means of the Joule effect, diminishing L 11
) and the THERMAL EFFICIENCY η c =
(l-λ c
) =
Q/Q R
(which disperses HEAT by means of Gas leakage and THERMAL-CONDUCTION, increasing Qi 1
). In practice we may calculate the EFFICIENCY η=L/Q of the IDEAL-CYCLE (Otto, Diesel, Joule, etc.) while the REAL-EFFICIENCY η R
=L R
/Q R
may be directly measured on the Drive Shaft or obtained by multiplying η=L/Q by ηo =
ηcη A>
a corrective factor that includes all the FRICTIONS (η 0
), including those THERMAL
η R = ηη 0 = ηη c η A = η(l - λ c )(l - λ A ) = ^x -^-x^- = ^ (1)
V VR L VR
Practically, for all HEAT-ENGINES the losses of WORK are on average 6% with η A =0.94 and 9% those of HEAT with η c =0.91, as confirmed by the Indicated-Cycles (3 0 ) where η o =ηcη A reduces by about (6+9)=15% the Efficiencies η=L/Q of the respective Ideal-Cycles. Thus (1) can assume the following approximated Statistic value: η R = ηη 0 = ηη c η A = L R /Q R = (l -0λ5)η = 0.85η (2)
In effect this result (η R
=0,85η) approximates all ROTATIONAL-ENGINES, in particular TURBINE ENGINES, but not RECIPROCATING-ENGINES (Explosion, Diesel) where instead we meet a value significantly INFERIOR, as indeed results from a comparison with the EFFECTIVE-REAL-EFFICIENCY
CRITERIA OF OPTIMIZATION
In the System O(p,V) one of the two Cycles inverts the Isentropic Curves with respect to the other, (pV k =C)*→(Vp k =C), therefore they are geometrically Identical and carry out the same Work (L≡130 Joules), which should have the Real-Efficiency (2).
However this result η R- .0.85x0.448--0.38 is true for the Turbine Engine but not for the OTTO ENGINE where in place of η R ≡-0.85x0.565≡0.48 we obtain approximately an average value of η R ≤(0.48+0.38)/2=0.43, with a DEFICIENCY of 24%, (0.43/0.565)=( 1-0.24). Nonetheless, apart from these hypothetical approximations, the sums just do not add up. Generally it is supposed that in Reciprocating-Engines the drop in Efficiency (≤24%) depends not only on the Frictions λ=λ c λ A (≡15%) but also (in agreement with CARNOT's Theorem) on the so-called Organic-Efficiency (η O --9%) of Combustion, involving a series of Chemical Reactions (not so very convincing), as one might try to balance a financial account, and it is not taken into account that that decrease (≤9%) is already included in (2) and (3), in as much as IDEAL-CYCLES (Otto, Joule, etc.) ignore Combustion (as if it did not exist). Excluding the Organic-Efficiency (η 0 ) it remains to be understood what the TRUE-REASONS for the DISCREPANCY are. In the absence of alternatives, apart from the Frictions (λcλ A ), other CAUSES which may produce losses in Efficiency depend necessarily on mechanical DIFFERENCES between the ENGINES, that is the set of mechanisms that connect the COMBUSTION-CHAMBER to the DRIVE SHAFT, almost inexistent in TURBINES but significant and well-defined in RECIPROCATING-ENGINES, actioned by the CRANK- CONNECTING-ROD-MECHANISM (fig. 5). A confirmation emerges from the fact that in Turbines the Thermal Forces act around the Drive Shaft, whereas in Reciprocating-Engines those Forces are concentrated at DEAD- POINTS, when the Mechanism is aligned, where they are not useful.
In effect, in normal OPERATING modes (after the engine has started to run), every Internal Combustion HEAT- ENGINE SELF-REGULATES in REAL-TIME its own ENERGETIC-BALANCE, that is the THERMAL (δQ) and MECHANICAL (δL) exchange with the Exterior. In other terms the ENGINE runs itself, AUTONOMOUSLY, INDEPENDENTLY, as a "CLOSED-BOX", therefore it is indeed its own MECHANISM that determines all the Thermodynamic Conditions of the MOVEMENT, including the EFFICIENCY.
CRITERIA OF OPTIMIZATION
In essence, with the sum of the FRICTIONS (λ=λ c
λ A
) and the corrections (1) and (2), we can therefore say that the MECHANISM produces two RECIPROCATING-IDEAL-CYCLES (3)<→(3*), the PRIMARY (3) in the Combustion Chamber and the INDUCED (3*) around the Drive Shaft, as can be seen in figures 12-19. We can also add that, by means of the Mechanism, the PRIMARY-CYCLE (3) occurring in the Combustion Chamber produces the INDUCED-CYCLE (3*) around the Drive Shaft, and vice- versa. In the Systems O(F,s) and ω(T,S) (figs. 12 & 13) these two RECIPROCATING-CYCLES (3<→3*) certainly have the same AREA but different GEOMETRIC-FORMS, therefore they produce the same WORK (L=L*) with EFFICIENCIES generally different (ηe3)≠(η*e3*). However only the INDUCED-CYCLE (3*) actions the DRIVE SHAFT, therefore it could (ideally) be located in a hypothetical Toroidal-Cylinder around the Drive Shaft (as in Turbines) ignoring the rest of the Engine (as if it did not exist). In doing so the Thermodynamic-Scheme of the ENGINE would be significantly simplified, taking into account that (in every case) in (1) and (2) we must substitute the INDUCED- CYCLE-Efficiency η*e(3*) for the PRIMARY ηe(3), in order to obtain the following EFFECTIVE-REAL- EFFICIENCY (η E
), valid for every HEAT-ENGINE:
Therefore we need to trace the curves (figs. 12 & 13) for the two Reciprocating-Ideal-Cycles (3*→3*) in order to know the EFFICIENCY (η*) of the INDUCED-CYCLE (3*), that we will try to OPTIMIZE until we obtain the Maximum Value (η*) max modifying opportunely the POWER-TRAIN-ASSEMBLY, being the Mechanism which connects the Combustion Chamber with the Drive Shaft. In such a way, the EFFICIENCIES of RECIPROC ATING- ENGINES may be increased by 20-30%, in certain cases exceeding those (η=L/Q) of the PRIMARY-CYCLE (3) occurring in the Cylinder, unlike with Turbines where equations (2) and (4) approximately give (η≤η*). For this reason we shall limit ourselves here to the OPTIMIZATION of the EFFICIENCY of Reciprocating-Engines only (Explosion, Diesel), that is the Eccentric-Crank-Connecting-Rod-MECHANISM (fig. 6) where the Drive Shaft is translated by the OPTIMAL distance h=|OH|>0 with respect to the Axis (PH) of the Cylinder, in the direction of positive rotation <p=(AOP*). In order to start let us remember that the Physical State of a Mass M 0 (kg) of Gas (Ideal, Real) depends only (Gibbs) on two VARIABLES. Assigning the pair p=F/ A(Pa) and V=As(m 3 ) with the other pair F=Ap(N) and s=V/A(m), bound by the Area-Constant A (m 2 ), taking into consideration the two Physical-Constants
CRITERIA OF OPTIMIZATION
R(J/K) and k=(C p
/c v
), thus the Physical-Temperature RT(J), the Physical-Entropy δ(S/R) (dimensionless) and the Internal-Energy δU(J) all assume the following expressions in finite terms:
The Crank-Connecting-Rod-Mechanism (figs. 5 & 6) and the equations (5) consent the tracing of curves figs. 12 and 13 in the Systems 0(F, s) and ω(RT, S/R) of the two RECIPROCATING-CYCLES (3)<→(3*) of the RECIPROCATING-Engines, the PRIMARY (3) occurring in the Combustion Chamber (dashed lines) and the INDUCED (3*) acting around the Drive Shaft (continuous lines).
Moreover, (figs. 3 & 4) whichever pair of Reciprocating-Cycles (3<→3*) becomes more explicit in the Mechanical Systems O(F, s) and O(F*,s*) where the two Isentropes (S/R) and (S/R)* create the four Points (X<-»X*) and (Y<→υ*) of Thermal Inversion (δQ=O), that is the THERMAL-INCREASE Q 0 which conditions the Energetic- Balance and the Efficiency of the INDUCED-Cycle:
—J Qi = Q YAX = Uy x + L YAX I Q x = Q YAX - U γx + L YAX
[Q 2 = Q YBX = U Y x + L YBX [Q * = Q * Bχ - U γx + L * YBX
j(Q; -Q 1 ) = (Q; -Q 2 ) = iir[(FX -F y Y y ) -(F χ8χ -F y s y )]= Q 0
I n = LZQ 1 => η * = L/Q; = L/(Q, + Q 0 )
The novelty of equations (5), (6) and (7) and others of this type, that for the sake of brevity will shall omit here, consists in the fact that they separate the THERMAL-Magnitudes (of the first members) from the MECHANICAL- Magnitudes (of the second members). This is sufficient to DEMONSTRATE that the ENTROPY dS=δQ/T cannot represent CARNOT's Theorem, the MEASURE of the Thermodynamic-Irreversibility of CLAUSIUS, or the common State-Function (5) S=f(p,V)=g(F,s) which has always been considered the UNIQUE Mathematical expression of the SECOND-LA W-OF-THERMOD YN AMICS. At this point we consider that the above (Title, Summary and Description) should suffice in order to qualify the Characteristics and Aim of the research, reserving the right to present further indications (on request). However, for greater clarity, it should be wise to add some DETAILS accompanied by Figures, Tables, Diagrams, Numerical examples, which we will try to describe briefly, taking into account possible MODIFICATIONS and possible ERRORS, omitting for the sake of brevity to comment on the processes that lead to the Thermodynamic Equations (5), (6) and (7) and those that will follow.
CRITERIA OF OPTIMIZATION
1) IDEAL-CYCLES. The Transformations occur in a CYLINDER of Section A(m 2 ), where the Force F(N) creates the Pressure p=F/A(N/m 2 ) against the piston that runs the distance s(m) of Volume V=sA(m 3 ). Of significant importance are the PHYSICAL-TEMPERATURE RT(J) and the PHYSICAL-ENTROPY S/R(dimensionless), which depend (5) on k=C p /c v (dimensionless) but ignore the other Constant R(J/K) and are thus valid for all FLUID- THERMODYNAMIC systems. In the new Mechanical O(F,s) and Entropic ω(RT,S/R) Systems the Areas (δL) and (δQ) represent the same quantity of ENERGY (Joules).
2) CENTERED-CRANK-CONNECTING-ROD-MECHANISM (Fig.5) represents the POWER-TRAIN- ASSEMBLY of the current RECIPROCATING-HEAT-ENGINES, where the Connecting Rod is assigned b=|PP*|, the Crank r=|OP*|, the Piston Stroke c=2r=|PiP 2 |, the clearance pocket of the combustion chamber c o =|PoPi| and the useful length of the Cylinder c,=(c o +c). The MOVEMENT depends on the only Independent Variable φ=(AOP*), when at the extremes (P 5 P*) of the Connecting Rod (PP*) the Thermal Forces (F,F*) produce (in every instant) the movements (ds,ds*). Consequently the following Equations compare the Basic-Work (δL=Fds=F*ds*=δL*=F*) done to the Forces (F,F*) in the Points (P,P*), that Vφ≥O defines as the TRANSMISSION-RATIO ε(φ) between the Forces (F*/F) and the movements (ds/ds*): fδL = Fds = F * ds * = δL *
< v H ε(φ) = A = i: w
I ds F
3) ECCENTRIC-CRANK-CONNECTING-ROD-MECHANISM (figs. 6 and 7). Identical to the above, except that the Drive Shaft is TRANSLATED a distance h=|OH|>0 from the axis of the Cylinder. In this case, assigning also the Extension λ=b/r>2 and the Eccentricity p=h/r>0 of the Mechanism, the following EQUATIONS define the extreme Inclinations (α,),(α 2 ) of the Connecting Rod, the ratio Stroke/Radius (c/r), the Compression Phases (A— >B) and the Expansion (C-→D), the movement of the piston s(φ)= |P 0 P|, and finally the relationships that optimize the Transmission Ratio ε(φ):
CRITERIA OF OPTIMIZATION
These EQUATIONS may be easily derived by paying a little attention to the figures. Even if (together with the others) they are not indispensable for the aims of the Patent, however, they become useful in the dimensioning and OPTIMIZATION of the ENGINES, especially in the compilation of TABLES and the comparison of respective EFFICIENCIES. In order to complete the compilation of OTTO-DIESEL engines in the Systems O(F,s), ω(RT,S/R), it suffices to know the two extreme ISENTROPES (S/R)=constant, equivalent (5) to (Fs k
)=constant, obtained assigning (figs. 8 & 9) the coordinates of two Points, for example A(F=400N, s=0.08m) and D(F=900N, s=0.08m), that we will apply in the following:
4) OPTIMIZATION of the ECCENTRIC-MECHANISM (figs. 6 & 7). The OPTIMIZATION is obtained increasing as far as possible the PHASE-OF-EXP ANSION (π+αi)<φ≤(2π+α 2
), obviously to the detriment of the COMPRESSION α 2
≤φ<(π+αi), that is with the MAXIMUM value of sinα 2 =
p/(λ-l)<l. However, in order to maintain the continuity of the movement Vφ≥O it is wise to assume the Extension λ=b/r and the Eccentricity p=h/r in the intervals λ=b/r>2, p=h/r<(λ-l), including (fig. 7) the following OPTIMIZATION-LIMIT (λ=2) and (ρ=0.866<l), that we will apply only as an example:
5) IDEAL-OTTO (figs. 8 & 9). In a Cylinder of length C 1 =SCm, we assign the clearance pocket of the combustion chamber C 0 = lcm, the Compression Ratio δ=8/l=8, the Piston Stroke c=7cm, the Starting Compression Force F A =400N, the Final Expansion Force F D =900N, the Constant k=1.40. Then, assigning the Extension λ=b/r=4, with (9), (10) and (11) we can trace (Table A and figs. 12 & 13) the Otto in the Systems O(F, s) and ω(RT, S/R), of which we know the Isentropic of Compression (S/R) AB =6.1386, of Expansion (S/R) CD s8.1659, and the Coordinates of the four vertices (A, B, C, D).
CRITERIA OF OPTIMIZATION
The Enclosed-Areas represent ENERGY (J), that is WORK in the System O(F,s) and HEAT in the Plane ω(RT, S/R), measurable (for verification) in the unique Scale: 1 cm 2 =10 Joules. Finally (5) and (6) consent us to calculate the Energetic-Balance of the OTTO, that is the Heat-Absorbed (Q 1 ), the Heat-Surrendered (Q 2 ), the Work L=(Qi-Q 2 ), where the Ideal-Efficiency (η=L/Qi) takes the same value (η≤0.565) of the example (3), as follows:
Q 1 = Q BC = U BC = ^(F C S C -F BSB ) = U(165λI - 73.52) = 229.73 J Q 2 = Q AD = U AD = ^(F D S D -F ASA ) = ^(900 -400) = 100 J
(13) L = (Q 1 -Q 2 ) ≡ 229.73 -100 = 129.73 J η = l -δ ι~k = L/Q λ ≡ 129.73/229.73 = 0.565
6) IDEAL-DIESEL-CYCLE (figs. 10 & 11). Referring to the Systems O(F,s) and ω(RT, S/R), the CYCLE occurs in the previous CYLINDER of Length Ci=8cm, clearance pocket of the combustion chamber c o =S B =O.5cm, Compression Ratio 5=8/0.5=16, Stroke c=7.5cm, Extension λ=b/r=4, with the values of F A =400N, F D =900N, (S/R) A -=6.1386, (S/R)D≥8.1659, with s C --0.8923, thus after having completed the 4 vertices (A, B, C, D) the equation (6)i is applied with the following result:
Q 1 = Q BC = (U BC + Z BC ) = ^(173 - 97) + 194.01(0.8923 - 0.5)≤ 266.11 J
Qi = Q AO = U AD = τh(F D s D - F A s A ) = ^(900 - 400)= 100 J
L = (Q - β 2
) ≡ (266.11 - 100) = 166.1 I J (14)
7) CENTERED-OTTO-ENGINE (Table A and figs. 5. 12 & 13). The CYCLE occurs in the CYLINDER of useful length Ci=8cm, assigning the clearance space of the combustion chamber c o =lcm and thus the Compression Ratio δ=8/l=8, while the ENGINE is actioned by the Centered-Mechanism (fig.5) p=h/r=0, of Extension λ=b/r=4 and Crank r=3.5cm. From (9) and (10) we obtain the Transmission Ratio ε(φ)=F*/F=ds/ds*, the Compression O≤φ≤π and Expansion π≤φ≤2π Phases, as well as the two reciprocating CYCLES (3)<→(3*) in the Planes O(F,s) and ω(RT,S/R), the PRIMARY (3) occurring in the Cylinder (dashed lines) and the INDUCED (3*) around the Drive Shaft (continuous lines), measured in the Scale: 1 cm 2 = 10 Joules. Finally (6) and (7) consent us to calculate the Thermal-Increase Q 0 between (3)<→(3*) in the 4 Points (X<→Y) and (X*<→Y*) of Maximum-Entropy (SIK- S*/R) M where Table A is reduced to 2 lines of Compression (φ=l 13°) and Expansion (φ=247°), from which is
CRITERIA OF OPTIMIZATION
obtained the Thermal-Increase Q 0 and the Efficiency η* of the INDUCED-CYCLE (3*), yet to be OPTIMIZED, that substitutes the (η=0.565) of the PRIMARY-CYCLE (3) already noted by (13):
(Table A) rρ 0 ≤ - 5 L -[(118.31 - 52.58) - (100.12 - 44.50)]s 25.28 J Q * = (Qx + Q 0 ) = (229.73 + 25.28) = 255.01 J (15)
[77 * = L/Q; = 129.73/255.01 ≤ 0.509
The fall (4) η E s(0.85x0.509)s0.43 in Efficiency (η*=0.509) of the INDUCED-CYCLE is thus confirmed.
8) ECCENTRIC-OTTO-ENGINE ( " Table B and figs. 7. 14 & 15). To the ECCENTRIC-MECHANISM (fig.7) with h=|OH|>0, we apply the OPTIMIZATION-LIMIT (12), that is the Extension λ=b/r=2 and the Eccentricity p=h/r=V3/2, for the same CYLINDER, c o =lcm, c,=8cm, δ=8. Thus from (9) and (10) we obtain the new Crank arrangement rs2.95cm, the new Phases of Compression 60°<φ<197° and Expansion 197°<φ<420°, and finally the two Reciprocating Cycles (3)<-»(3*) in the Mechanical O(F,s) and Entropic ω(RT,S/R) systems, with the scale: lcm 2 = 10 Joules.
In this case the four points (X<→Y), (X*<→Y*) of Maximum Entropy can be found in the two lines (φ=133°) and
(φ=262°) of Table B, where the equations (6) and (7) define the Thermal Increase Q 0 and the Efficiency (η*=0.634) of the new OPTIMIZED-INDUCED-CYCLE (3*), that must substitute (η*=0.509) obtained using equations (15):
(Table B)
Q 0 s ^ j - [(110.13 - 58.54) - (104.21 - 42.60)] = -25.05 J • Q * = (Q 1 + Q 0 ) = (229.73 - 25.05) = 204.68 J (16) η * = L/Q; ≤ 129.73/204.68 ≤ 0.634
The Optimization-Limit (12) therefore produces an Induced-Efficiency η*=0.634 greater (0.634/0.509=1+0.246) by about 25% than the Efficiency (η*=0.509) of the Centered-Engine, even exceeding (0.634/0.565=1+12) by 12% the Efficiency (η≤0.565) of the Primary-Cycle occurring in the Cylinder. This means that in the two intervals λ>2 and 0<p<(λ-l) there must be a Equilibrated-Optimization Q 0 =O with η=η*, being η≥η*, VQ 0 >0, or η≤η*, VQ o <0.
CRITERIA OF OPTIMIZATION
9) CENTERED-DIESEL-ENGINE (Table C and figs. 5. 16 & 17). The MECHANISM of Extension λ=b/r=4, and the CYLINDER of length c,=8cm, are the same as the OTTO-ENGINE, where however the clearance pocket of the combustion chamber is assigned c o =O.5cm, the Compression Ratio 8=8/0.5=16, the new Crank arrangement r=7.5/2=3.75cm. Then (9) and (10) define the two Reciprocating-Cycles (3)<→(3*) in the Systems O(F, s), ω(RT,
S/R), the PRIMARY CYCLE (3) occurring in the Cylinder (dashed lines) and the INDUCED CYCLE (3*) occurring around the Drive Shaft (continuous lines), in the Scale: 1 cm 2 =10 Joules.
In this case in the two lines (φ=127°) and (φ=233°) of Table C the four points of Maximum-Entropy (X<→Y) and
(X*<→Y*) converge, and define the Thermal-Increase Q 0 and thus also the Energetic-Balance (6) and (7) of the new
INDUCED-DIESEL-CYCLE (3*) together with the EFFICIENCY (η*):
(Table C)
[Qo = 0 * 4 -065.20 - 73.42) - (118.64 - 52.73)] ≤ 64.. ( 675 J Qi = (Qi + Q 0 ) = (266.11 + 64.675) = 330.785 J (17) [77 * = L/ Q * = 166.11/330.785 = 0.502
The Induced-Efficiency (η*=502) comes out inferior by about 20% to that of the Primary-Cycle of (14) (η.=0,624). Also in this case the approximation (4) η E s0.85x0.502__0.43 becomes reliable by ignoring the Combustion. 10) ECCENTRIC-DIESEL-ENGINE (Table D and figs. 10 & 11). To the previous Centered-Engine (14) we apply the OPTIMIZATION-LIMIT (12), that is the Extension λ=b/r=2 and the Eccentricity p=h/r=V3/2. Using equations (9) and (10) we thus calculate the new Crank arrangement rs2.95cm, the Phases of Compression 60°<φ<197° and Expansion 197°<φ<420°, as well as the two Reciprocating-Cycles (3) and (3*), these being tabulated and drawn in the Systems O(F, s) and ω(RT, S/R). Also in this case we should highlight the four points (X, X*) and (Y, Y*) of Maximum Entropy (S*/R) M whose coordinates are in the two lines (φ=145°) and (φ=248°) of Table D, where the equations (6) and (7) allow us to calculate the Thermal-Increase Q 0 (from 3 to 3*) and thus the Efficiency (η*) of the INDUCED-DIESEL-CYCLE (3*) acting around the Drive Shaft, that as before substitutes the Efficiency (η=0,624) of the PRIMARY-CYCLE (3) occurring in the Cylinder:
(Table D)
CRITERIA OF OPTIMIZATION
fβ o = ^[(157.77 - 78.37) - (123.49 - 50.15)] ≤ 15.15 J ' βi * = (Q + βo) ≥ (266.11 + 15.15) = 281.26 J (18) η * = L/Q * = 166.11/281.26 = 0.591
This Induced-Efficiency (η*--0.591) is inferior by 5% to that of the Primary-Cycle occurring in the Cylinder (3) of
(14) where (η=0.624), but is superior by 18% to the Induced-Efficiency of the Centered-Diesel-Engine (17) where
(η*≤0.502).
11) CONCLUSION. The EFFICIENCY of all the Internal-Combustion-Heat-Engines may be increased significantly by OPTIMIZING the POWER-TRAIN-ASSEMBLY, opportunely modifying the respective
MECHANISMS that transmit the Thermal-Forces of the Gas in the Combustion Chamber to the Drive Shaft.