Double piston engine

A double piston engine has a medial shaft between two pistons which reciprocate in opposed cylinders. From the pistons extend outer piston shafts which serve as control shafts. The outer ends of the cylinders are provided with inlet ports and control recesses while the control shafts have also control recesses and the meeting of the control recesses defines the inlet of the fluid into the cylinders. More details serve to combine a plurality of double piston engines to work in unison in timed relation, to increase the power per a given weight or to use the engine as a hydrofluid conveying combustion engine as well as the prevention of dead spaces by specific valves or configurations and locations. A piston may form a first piston portion and a plurality of secondary piston portions with the sum of the cross-sectional areas of the secondary piston portions equal to the cross-sectional area of the first piston portion.

BACKGROUND OF THE INVENTION 
1. Field of the Invention 
This invention relates to piston engines and partially to double piston 
engines. Such double piston engines often operate as free piston engines. 
They may, however, also be provided with rotary means to control the timed 
relation of operation of the pistons in the cylinders. 
2. Description of the Prior Art 
A double piston engine is described in my U.S. patent application Ser. No. 
06-529,254. In said application means are provided between the pistons to 
transfer the power of the combustion engine cylinders into reciprocating 
pistons of hydraulic pumps. Thereby the engine works as a hydrofluid 
combustion engine. Similar engines of hydrofluid conveying combustion 
engines are known from my U.S. Pat. Nos. 3,174,432; 3,260,213 and 
3,269,321. A free piston engine is known from U.S. Pat. No. 4,385,597 to 
Frank Stelzer. The mentioned patents serve specific purposes and obtain 
them partially or totally. However, all of them are either still too heavy 
to permit the application in vertically taking off aircraft or they fail 
to have enough uniformity of flow if they are used to supply a flow or 
flows of hydraulic pressure fluid. Some of the mentioned engines also fail 
to have a uniform supply of power. In my mentioned earlier patents the 
forces of the combustion engine pistons are in equilibrium with the force 
consumption of the pistons of the hydraulic pumps. However, such 
equilibrium goes on the expense of uniformity of supply of power over 
time. The hydraulic hoses and pipes broke, thereby, under ununiform 
deliveries of fluid. 
SUMMARY OF THE INVENTION 
It is the object of this invention to increase the power of an engine per a 
unit of weight. 
Another object of the invention is to provide a combustion engine with 
simple inlet and outlet means. 
A further object of the invention is to provide a double piston of little 
weight in order to permit higher RPM of the engine. 
Still another object of the invention is to run a plurality of double 
piston engines in timed relation relative to each other and to provide the 
means thereto by little weight of the components. 
Still a further object of the invention is to provide a little weight 
powerful aircraft engine. 
A still further object of the invention is to provide a flow of fluid or 
plural flows of fluid out of the engine with an almost uniform flow. 
Other objects of the invention are dead space preventing valve means, inlet 
recesses, control shafts, control recesses and other inlet or outlet 
means. 
More objects of the invention will become apparent from the description of 
the preferred embodiments and from the appended claims. The mentioned 
claims thereby serve partially also as the description of the aims and 
objects of the invention as well as a description in part of the preferred 
embodiments of the invention.

DESCRIPTION OF THE PREFERRED EMBODIMENTS 
FIG. 1 shows a sectional view of my mentioned earlier patents. It has a 
cylinder 2 with a therein reciprocating piston 4 which periodically varies 
the volume of the chamber 1. The piston shaft 777 pumps hydraulic fluid in 
chamber 111 when the engine cylinder 4 reciprocates. The fluid is 
delivered through an exit valve by which the arrangement becomes a 
hydrofluid conveying combustion engine. 
FIG. 2 illustrates in a longitudinal sectional view the engine of the 
mentioned patent to Frank Stelzer. It has between the engine pistons a 
medial pre-charging piston with respective inlet and transfer means. 
FIG. 4 is a longitudinal sectional view through another hydrofluid 
conveying combustion engine of my mentioned earlier patents. Engine piston 
4 has an interior chamber 21 into which a stationary bar 25 sealingly 
extends. Bar 25 has passages with an entrance valve 22 and an exit valve 
23. When the engine piston moves upwards in the compressioin stroke, the 
hydraulic fluid enters over valve 22 into the interior piston chamber 21. 
At the expansion stroke of the engine piston the valve 22 closes and the 
hydraulic fluid is pressed out over valve 23. Thereby the expansion or 
power stroke of the engine piston 4 supplies a flow of hydraulic fluid out 
of valve 23. The engine power is transformed into hydraulic fluid power. 
However, the flow of hydraulic fluid is not uniform over time. It is 
therefore important in accordance with the present invention, to find a 
way of calculating the actual appearances. 
FIG. 3 and FIG. 5, therefore, illustrate the basic principle of the engine 
in a schematic with the definition of the geometrical and mathematical 
values. In FIG. 3 piston 4 has a shaft 7 only in one axial direction, 
while in FIG. 5 the engine piston 4 has two shafts 7, each one in each of 
the axial directions. The shafts 7 extend through the covers 8 or 8 and 3 
respectively. The exhaust passage is shown by 6 and the maximum of piston 
stroke with compression or expansion is obtained when the top face of 
piston 4, the face 5 opens the exhaust passage 6. 
The actual stroke of the piston at compression starts by "H1" and is 
defined to be: "H". Passage 9 is provided to prevent a compression of 
fluid below the bottom of piston 4. The radius and the diameter of the 
piston are shown by "R" and "D" respectively. 
For the compression actual pressure shown in the diagram of FIG. 6. How 
these pressures and other values are found will be shown in the analysis 
of the engine. 
ANALYSIS OF THE ENGINE 
For the compression or expansion of the gas in the cylinder off the engine 
the basic gas equation (1) applies. 
EQU P.times.V.sup..eta. =constant (1) 
Therein "P" is the pressure, "V" is the volume and ".eta." is the adiabatic 
exponent. It follows: 
EQU P.sub.1 .times.V.sub.1.sup..eta. =P.sub.2 .times.V.sub.2.sup..kappa.(2) 
and: 
EQU P.sub.2 =P.sub.1 (V.sub.1 /V.sub.2).sup.n (3) 
and: 
EQU P.sub.2 =P.sub.1 V.sub.1.sup..eta. V.sub.2.sup.-.kappa. (4), 
since 1/V2 high "n" is V2 high minus n. In the following ".kappa." will be 
substitutet for "n" for the ease of typing. This exponent "n" is between 
1.3 and 1.42. 
The cross sectional area of the piston 4 is defined by equation (5) as: 
EQU F=R.sup.2 .pi. or F=D.sup.2.pi. /4 (5) 
with "D"=2R and "pi"=3.14. 
In the sample of the analysis the maxium of stroke will be 10 cm and the 
cross sectional area of the piston 4 will be 100 squarecentimeter by which 
the cylinder chamber 1 will have a volume of 1000 CC with 
CC=cubiccentimeter. For the cylinder of FIG. 5 with the piston shaft of 
diameter "d" and the piston of diameter "D" follows equation (6): 
##EQU1## 
It would be helpful if equation (4) could be transformed to get "H" as the 
variable. At first impression such equation looks to read: 
EQU P.sub.2 =P.sub.1 (.sup.2.pi. /4).sup..kappa. H.sub.1.sup..kappa. 
H.sub.2.sup.-.kappa. (7) 
however, such equation would be wrong, because equation (3) would bring 
equation (8) as follows: 
##EQU2## 
wherein "F" appears above and below the fraction line. That simplifies the 
equation to equation (9) as follows: 
EQU P.sub.2 =P.sub.1 H.sub.1.sup..kappa. H.sub.2.sup.-.kappa. (9); 
in which "P1" and "H1 high minus n" are constants. The variable now is the 
actual stroke "H"="H2". 
One will find that the compression pressure may become very high and the 
compression pressure is shown in FIG. 6 over the actual piston stroke. The 
calculation is done until a distance of 1 mm of the top face 5 of piston 4 
from the cover 3. The compression pressure would then be already about 500 
atmospheres and if the combustion occurs with air ratio "lombda"=1, the 
combustion pressure would reach about 2000 atmospheres. This brings to 
light that at such high compression ratios the walls of the cylinders 
would break because they can not withstand such high internal pressures. 
It is now convenient to find the medial pressure at compression or 
expansion because having it the power is simply this medial pressure "p" 
multiplied by the stroke "H". In the following the respective equation(s) 
will be developed: 
##EQU3## 
Equation (11) follows from equation (10) since it is known from equation 
(8) that the areas eliminate. 
From these equations it is easy to find the actual work (work "A") by 
multiplying the medial pressure with the area "F" and the stroke "H". The 
stroke difference then eliminates and it follows: 
##EQU4## 
FIG. 8 illustrates the known P-V diagram, however, for the values which are 
applied in this analysis. Introducing the indices "c" for compression, "e" 
for expansion; one obtains: 
##EQU5## 
The values which are obtained, are, as defined, works but not powers. To 
obtain the power therefrom, the work would have to be multiplied with the 
number of strokes per second. Remember; 
EQU work=Kgcm; power=Kgcm/sec. 
The engine of the before mentioned patent to Frank Stelzer of the former 
art is called the "STELZER ENGINE" and in literature about the Stelzer 
Engine in magazines and newspapers in German, French and Japanese 
languages it is reported, that the Stelzer engine with a piston of 5 Kg 
makes 30,000 double strokes per minute. This requires a further 
investigation. 
To evaluate the maximum of possible strokes of a free piston engine, 
Newtons law of force is brought to attention, which defines: (with 
force=K) 
##EQU6## 
and with Newton: 
EQU k=mb (18) 
with: 
V=velocity; K=force=Kg 
t=time (seconds) 
H=way, stroke (meters) 
b=acceleration (m/sec square) and: 
m=mass=weight of piston/9.81 m/sec square,gravity) 
Therefrom follows the acceleration of the piston of the free piston engine 
as follows: 
##EQU7## 
and equation (17) can be transformed to: 
##EQU8## 
Therein "b" may be inserted from equation (19) to obtain: 
##EQU9## 
and for the force "K" the value K=F.times.P may be inserted to obtain the 
basic acceleration equation for the free piston engine as: 
##EQU10## 
The number of strokes per second is obtained by multiplying with 1/t and, 
consequently, is: 
EQU EH/s=1/t (24) 
wherein "E" indicates a single one way stroke and for "H" the difference 
(H1-H2) might be inserted. 
Therefrom follows: 
##EQU11## 
with the constant: "B": 
EQU B=8/d.sup.2 .pi. (26) 
Just for memory, the constant "B" could further be shortened as follows: 
##EQU12## 
and the number of strokes (one way strokes) per second would be: 
##EQU13## 
One now has a beautiful equation for the calculation of the number of 
strokes which are maximally possible, but it will be seen soon that it is 
not so easily possible to calculate with it. That will become apparent at 
hand of the inquiry about the announced number of strokes of the Stelzer 
engine. 
Neglecting accuracy and assuming at first glance that the piston would be 
accelerated by the maximum of pressure at the combustion at eta=40 
(eta=compression ratio) one would obtain with n=1.35 (in all further 
calculations n shall all times be 1.35 in this analysis): 
EQU P.sub.2 =P.sub.1 H.sub.1.sup..kappa. H.sub.2.sup.-.kappa. =1.100.sup.n 
.multidot.2.5.sup.-.kappa. 
=1.501.multidot.0.29026=145.42BAR=1454200KG/m.sup.2 
The constant B therein brings B=200.04 with D in meters; the mass "m" of 
Stelzers motor was announced to be 5 Kg; eta=40 gives H2=0.25 mm and tyhe 
number of strokes per second would then be: 
EQU EH/s=[200.04.times.0.5.times.0.0975.times.1454200.sup.-1 ].sup.-0.5 
=[200.04.multidot.0.5.multidot.0.0975.multidot.6.877.sup.-7 ].sup.-0.5 
=[6.6996.sup.-6 ].sup.-0.5 =386EH/s 
which corresponds to 386.times.30=11 588 DH/min with DH=double strokes per 
minute. This calculation was done for the compression stroke. For 
combustion at lomda=1 the expansion pressure would be four times higher, 
145.45 bar .times.4=581.68 bar inserted, would yield: 
EQU EH/s=[200.04.times.0.5.times.0.0975.times.5816800.sup.-1 ].sup.-0.5 
=772EH/s=23169DH/min. 
Therefrom the compression pressure would have to be subtracted, but one 
could at this first glance get the impression that the Stelzer engine in 
case of extremity of luck could make the announced 30,000 double strokes 
per minute. For calculating the strokes from the compression calculation 
for lombda=1,(four times higher pressure at expansion stroke), the result 
of the calculation for the compression stroke would have to be multiplied 
with .sqroot.(4-1)=1.73. The number of double stokes per minute would then 
be 11 588.times.1.73=20 047 DH/min. 
The above calculation was done, however, at a first glance only with the 
wrong assumption that the maximum of pressure would act over the entire 
expansion and compression stroke. That is, however, not the case because 
the pressure drops immediately when the piston moves away from the 
combustion point (the inner dead point at 2.5 mm) towards the outer dead 
point, the exhaust location of the piston. For a next simplified 
consideration it might be assumed that the arithmetic medial pressure of 
the stroke might be inserted. Neglecting the compression stroke, the 
arithmetic mean pressure at the expansion stroke would be 
Pme=(P6+P4)/2=(582+4)/2=293 bar. The calculation with equation (29) would 
bring: 
EQU EH/s=[200.04.multidot.0.5.multidot.0.0975.multidot.2930000.sup.-1 
].sup.-0.5 =[3.328.sup.-6 ].sup.-0.5 =548EH/s=16444DH/min 
Considering the subtraction of the compression pressure with 
.sqroot.3/4=0.866, this value multiplied with the 16,444 DH/min gives 
14,240 double strokes per minute =14,240 DH/min. The maximally possible 
number of strokes has already drasticly reduced at this slightly more 
accurate calculation. 
The above consideration is, however, also only a very simplified and wrong 
consideration. If one looks at the P-V diagram of FIG. 8 one sees that the 
curves of the pressures at compression and at expansion are no straight 
lines but curves. The next still only slightly more accurate assumption 
might now be to use the medial pressures of the compression and expansion 
strokes from equation (11). Inserting these values one obtains: 
##EQU14## 
and the strokes per second and double strokes per minutes would be: 
EQU EH/s=[200.04.multidot.0.5.multidot.77269.sup.-1 ].sup.-0.5 =[1.26.sup.-4 
].sup.-0.5 =89EH/s=2670DH/min..times.1.73=4619DH/min. 
The maximally possible number of strokes per unit of time have now really 
drasticly decreased. They are down to almost a fifth of the first 
calculation. However, even this consideration is not accurate, because 
equation (2) is valid only for a constant acceleration over the entirety 
of the way of stroke. In actuality in the free piston engine the 
acceleration varies at any moment of the stroke of the piston. The 
inventor of this application has tried since a long time to find an 
analytic mathematical formula for the actual acceleration of the piston of 
the free piston engine which would take into consideration the at all 
times varying acceleration during the stroke of the piston. Regrettably, 
however, such formula has not yet been found. The remaining possibility to 
increase the accuracy is, therefore, to use the medial pressure for small 
intervals of the stroke and insert them into equation (11). That is not so 
simple but it can be done if a respective form is used. Such suitable form 
is shown in FIG. 9 and in FIG. 10 the form of FIG. 9 is used to calculate 
the above example of values actually through. It is learned from it that 
the maximally possible number of strokes is still far less than the number 
of strokes of the last calculation there before. 
For further improvements of the consideration procedures the equation (29) 
is once looked upon again. It reads: 
EQU EH/s=[Bm(H.sub.1 -H.sub.2)p.sup.-1 ].sup.-0.5 (29) 
or written in the other form: 
##STR1## 
which could still written differently bu using the rules of calculations 
with powers and roots as follows: 
##STR2## 
From equation (30) is is immediately visible that the number of strokes 
increases with smaller values below the fraction line. 
Therefrom the following rules are obtained: 
1. The number of strokes increases with the root of decrease of the mass. 
2. The number of strokes decreases with the root of increase of the mass. 
3. The number of strokes increases with increase of the root of the medial 
pressure. 
4. The number of strokes decreases with the root of decrease of the medial 
pressure. 
5. The number of strokes increases with the root of the decrease 
(shortening) of the length of the stroke. 
6. The number of strokes decreases with the root of the increase 
(lengthening) with the length of the stroke. 
(The rules 5 and 6 are, however, in practical application not all times 
suitable since with the variation of the lengths of the strokes the 
pressures also variate. This has to be considered in cases of applications 
of rules 5 and 6.) 
Samples of calculations with these rules may be seen in West German patent 
publication DE-OS-33 41 718.0 published on May 30, 1984. 
The mentioned German publication contains also in detail explanations how 
by the above established rules the sample of the Stelzer engine could be 
considerably improved. 
The stepwise calculation by stroke intervals as done in FIGS. 99 and 10 
could be eliminated if the actually acting medial pressure " P " could be 
calculated. That is still not possible and a graphic methode might, 
therefore, be suitable. Before considering a graphical solution, some 
mathematical results of applicant's considerations shall be memorized. 
They do not yet lead to a mathematical solution but may be helpful for 
steps of calculations for which they are shown in the following: 
Medial integral pressure "P" at compression and expansion, calculated from 
the volumes: 
##STR3## 
Medial integral pressure "p" at compression and expansion calculated from 
the strokes: 
##STR4## 
Medial integral pressure "P.sub..DELTA. " at H2 minus interval .DELTA.H for 
compression and expansion: 
##STR5## 
Medial integral value ".epsilon." of the compression ratio ".epsilon.": 
##STR6## 
Differential of pressure "P.sub.2 " relative to the stroke: 
##STR7## 
Medial integral of the differential of pressure "P2" relative to stroke: 
##EQU15## 
Caluclation of the time "t" if a medial acting pressure " P " would be 
known: 
##EQU16## 
(This calcualtion is valid only if the acting medial pressure " P " would 
be known. Rgrettably, this acting medial pressure is not yet known.) 
Calculation of the time "t" if the calculation from the pressure "P2" would 
be possible (which regrettably is not possible): 
##EQU17## 
MEMO 
##STR8## 
Medial integral pressure "P.sub..DELTA. " by difference P2 minus " ": 
##EQU18## 
Calculation of time "t" if P2 would be constant over stroke: 
##EQU19## 
Calculation of time "t" if "P" would be constant over stroke 
##EQU20## 
Calculation of time "t" if "P.sub..DELTA. " would be constant over stroke: 
##EQU21## 
Since the acting medial pressure " P " has still not been found it shall 
now be defined for the sample of the Stelzer engine which was calculated 
herebefore, at hand of compression ratio ".epsilon.=40". It can be 
obtained by modifying column 34 of FIG. 10 to: " P ". It yields: 
EQU .SIGMA.t(column 38)=0.0558; (.SIGMA.t).sup.2 =3.1136.sup.-3 ; t.sup.2 
=2.DELTA.Hm/K=2.DELTA.Hm/P.sup.F 
and 
EQU P =2.DELTA.Hm/F(.SIGMA.t).sup.2 =2.multidot.0.0975 
M.multidot.0.5/100.multidot.3.1136.sup.-3 =0.313138 
This value of only 0.313138 bar (Kg/cm square) is, however, a great 
surprise. At the start of the stroke the pressure P2 or P4 was extremely 
high. At the earlier calculations the medial pressures at compression were 
still a number of atmospheres but now the acting medial pressure is only a 
fraction of an atmosphere. That is so because the high pressures act only 
at extremely short times during the strokes. 
Since the result is such a big surprise the matter shall now be further 
investigated. The equation for the calculation of the time "t" was: 
##STR9## 
and can be transformed to: 
EQU P =2(.DELTA.H)m/Ft.sup.2 (45) 
It now looks as if the searched for acting medial pressure " P " could be 
found by summarizing the found values of the intervals to calculate with 
them. If that would work a so found medial acting pressure might probably 
in future be used if written in a graph. The acting medial pressures " P " 
could then be taken from such a graph and be used for calculation in the 
earlier established formulas. For that purpose equation (45) would have to 
be written to define that the sum of the intervals of the times "t" have 
to be used. On so obtains: 
EQU P =2(H.sub.1 -H.sub.2)m/F(.SIGMA.t).sup.2 (46) 
The result is shown in FIG. 39 and it is calculated in the table of FIG. 
39. 
((Memo: for control of the consideration equation (29) may be applied with 
the obtained " P ". The control calculation would bring 658 DH/min for the 
entire engine. That is different from the above consideration, and, 
consequently, the above defined calculation for " P " may not yet be 
correct and should be considered as such, be used only with care. 
EQU EH/s=[Bm(H.sub.1 -H.sub.2) P .sup.-1 ].sup.-0.5 
=[200.04.multidot.0.5.multidot.0.0975/3145].sup.-0.5 
=12.67EH/s=380DH/min.)) 
COMISON WITH OTHER ENGINES 
Applicant's 1978 aircraft engine with 811 CC run with 10,000 RPM and gave 
120 HP. The weight of the conrod plus piston per cylinder was about 500 
grams. The mass was thereby only about 0.05. Compression ratio was 
".epsilon.=9" about. Using these values in the above equations for the 
free piston engine one would obtain: .phi. of piston=6.1 cm. Stroke=6.3 
cm. 6.1.sup.2 .pi.=116.89 cm.sup.2 ; B=8/116.89=0.068. M=0.05. 
.DELTA.H=6.3 cm. For .epsilon.=9 from FIG. 39 follows P =0.3210 
kg/cm.sup.2 .times.3 for entire engine=0.963. 
##EQU22## 
This comparison shows that the aircraft engine could have run only 2011 
RPM if the free piston engine equations would be used. But actually the 
engine run 10,000 RPM. This shows that the equations for the free piston 
engine can not be used for the engine with a crankshaft. In the above case 
the crankshaft of the aircraft engine had a weight of about 9.5 Kg. The 
engine had four pistons and about 6 Kg were located at half of the radius. 
This gives a mass of about 0.15 per piston's crankshaft counter weights. 
This mass did, however, not make just the stroke, but 1 times pi/2=1.57 
times of the stroke as rotary movement. The kinetical energy of the 
counterweights of the crankshaft was, therefore, (1.57) square=2.47 times 
of the kinetical energy of the reciprocating piston of the free piston 
engine. Since the mass of the conrod plus piston was only 0.05 the 
kinetical energy was (0.15/0.05)/2=7.4/2 times higher than the kinetical 
energy required to accelerate the piston and its conrod. 
One obtains the following important conclusion: 
The common engine with a crankshaft has counter weights which move a 1.57 
times longer way than the stroke of the free piston engine is and thereby 
the engine with a crankshaft has a permanently available kinetical energy 
at a given revolution which overcomes the required acceleration forces 
which are required to accelerate the conrod and the piston to the 
reciprocating stroke. The crankshaft engine has thereby an ability to 
obtain any desired RPM (until it breaks) while the free piston engine does 
not have such a bank of avialable kinetical energy and is forced to 
accelerate its piston by the pressure in the cylinder at each individual 
stroke. Thus, the free piston engine is limited in the number of strokes 
while the engine with a crankshaft can obtain any desired RPM until it 
breaks or until the ports are too small to bring or expel enough fluid. 
Since in the free piston engine the compression requires at least one 
fourth of the power of the expansion stroke and since the expansion stroke 
must drive the compression stroke, the free piston engines loses at least 
one fourth of the energy of its fuel for the operation with the 
compression stroke. 
This is an important consideration and shall therefore be more deeply 
inquired. 
For that purpose FIG. 10 has in column 37 the kinetical energy of the 
piston of the free piston engine. Column 42 gives therefrom the HP of the 
engine. To check column 37 of FIG. 10 equation (13), which is a pure 
thermodynamic equation, may be used. It gives: 
##EQU23## 
Compared therewith column 42 in FIG. 10 gives 54.35 Kgm. The results are 
not equal but not very much different. It shows that the actual results of 
FIG. 10 are not too much wrong for the first calculation attempt. 
Comparing consideration for the balance of the energies: 
FIG. 12 shows the conrod and the piston of the mentioned aircraft engine of 
1978 in a 1:1 scale. It corresponds to the 750 CC Honda motorbike engine 
of the seventies. 
FIG. 13 shows the mechanism of the crankshaft engine with the therein 
applying equations. The equations are partially simplified by neglecting 
values of small results. 
At one half of a revolution the kinetical energy for the acceleration of 
the piston and conrod is taken out of the crankshaft and at the next half 
revolution it is added to the crankshaft by which the crankshaft maintains 
its kinetical energy over the time. For acceleration and slow down of the 
RPM of the crankshaft engine more or reduced fuel energy is supplied by 
opening the throttle wider or by reducing it. 
Improvements of the free piston engine: 
Using the rules which were established above it will now be attempted to 
improve the free piston engine for a greater number of strokes per 
revolution. 
FIG. 14 shows an important embodiment of a free piston engine of the 
invention in a 1:1 scale in longitudinal sectional view. The improvement 
compared to the Stelzer engine is a reduced weight of the piston to about 
1.5 Kg in case of a piston of steel. The Figure has additional 
improvements. However, the reduction of weight of the piston brings 
according to the in this specification established rules a considerable 
and important increase in the number of strokes which are possible in a 
unit of time. The detailed calculations of the number of strokes etc. is 
not given in this specification. 
In FIG. 14 a charger (turbo) supplies pre compressed air or air-fuel 
mixture from inlet 9 over control recess arrangement 15 into the working 
chamber (cylinder) 1. Head cover 3 is mounted onto the wall 2 of the 
cylinder. Inclined faces 14 and 13 may be provided on the cover 3 and 
piston 4 to streamline the flow of air or gas. The gas leaves the chamber 
1 through outlets or exhaust ports 6 when the piston has the location as 
shown in the Figure. The chamber 1 is now flashed. From piston 4 extends 
in the axially outward direction the piston shaft or control shaft 7 which 
has the control recess 7 which opens and closes the inlet port 9 to and 
from the cylinder or chamber 1 at the up and down stroke (reciprocation) 
of the piston. Shaft 7 may be provided with a piston ring groove 154 to 
have therein the piston ring (seal ring) 153. 
In FIG. 15 the engine portion of FIG. 14 which may also be used in a 
crankshaft engine, is shown in a scale reduced to one third and mounted to 
form with a second opposing cylinder a free piston engine. The weight of 
the piston is about 3.8 Kg and the engine of FIG. 15 would as free piston 
engine obtain about two times the number of strokes compared to the 
earlier discussed Stelzer engine. Details of calculation are available in 
the mentioned German DE OS. The bottom of FIG. 15 shows the opposed 
cylinder, cover and piston with pre-indices 6. The bottom portion of the 
engine acts similar as the top portion, however, at opposed strokes and 
times. The pistons 4 and 64 are connected by the medial piston connecting 
portion 60. When one of the cylinders 1 or 61 acts in the expansion stroke 
the opposed cylinder 61 or 1 acts in the compression stroke. Ignition 
means and fuel injection means are not shown in the Figures of this 
specification because they ar known in the art. The engine of FIG. 15 and 
the similar embodiments of this specification are thereby one cycle 
engines because at every stroke the engine has a power stroke. Once the 
respective cylinder is flashed and filled with fresh air, the piston moves 
and closes the exhaust ports 6,66, whereby the compression begins and the 
ignition occurs when the respective piston 4,64 is close to the cover 3,63 
while thereafter the direction of movement of the piston(s) reverses and 
the power stroke begins until the respective piston opens at the end of 
the power stroke the exhaust port(s) 6,66 for the exhaust of the used 
gases. 
As a further speciality of this Figure an exhaust collecter 16 is mounted 
around the medial portion of the cylinders and the exhaust ports 6,66 port 
into the exhaust collection chamber housing 16. 
FIG. 16 is a cross sectional view through FIG. 15 along the arrowed line 
XVI--XVI of FIG. 15 and illustrates that instead of providing just an 
exhaust gas collection chamber 16 the arrangement may include exhaust 
chambers 16 and additionally therefrom separated cool fluid supply 
chambers 19 with cool fluid supply entrances 18. They will press cooling 
fluid into the space 59 between pistons 4 and 64 around medial connecting 
portion 60 to cool the neighboring parts. A passage or a plurality of 
passages may be provided through the medial piston connecting portion 60 
in order to lead the cooling fluid also through the hollow piston shafts 
7,67. These passages are not shown in FIG. 15. Passages 20 may also be 
provided in the cylinder wall to connect with the free outside if so 
desired. 
FIG. 18 shows how the number of strokes per a given unit of time can become 
increased in accordance with the analysis of this specification. The top 
and bottom portions of FIGS. 14 or 15 are assembled to a medial housing 
57. In this housing 57 a crankshaft 54 is revolvably borne in bearings 56 
and it has the counter weights 52. Connecting rods 55 are borne by the 
crankshaft at 54 and connect to the piston(s) at 58. Cooling ribs 53 may 
be provided on the pistons. Now the formerly free piston engine has 
obtained a revolving crankshaft with the revolving mass which forms the 
bank for the containment and supply of the kinetical energy to accelerate 
the pistons to their reciprocating strokes. The number of strokes per unit 
of time of the engine of the invention of FIG. 17 can now make any desired 
strokes per unit of time until it would break. The limitations to number 
of strokes of free piston=double piston engines is now overcome by this 
Figure. Instead of using the term "connecting rod" for part 55 the common 
term "conrod" is used in this specification. 
FIG. 18 shows the velocity, acceleration and required forces K for the 
acceleration for reciprocation of piston and conrod of a sample of an 
engine over the rotary angle "alpha" of the crankshaft. 
FIG. 19 shows a diagram of the powers obtainable from a sample of an engine 
at different strokes and compression ratios. 
FIG. 20 shows in a cross sectional view through the housing 80, which 
brings longitudinal sectional views through the cylinder and piston 
arrangements, a multiple double piston engine of the invention. The 
housing 80 bears in 56 a crankshaft with an eccentric bearing portion 54 
which bears the conrods 46 to 48. The outer ends of the conrods connect to 
the double pistons at 43. This engine has 3 double cylinders 31 angularly 
spaced by 60 degrees. The engine might have any other number of double 
cylinders if they are respectively angularly spaced. Each cylinder 32 has 
two cylinder chambers 31 and 41 which are separated from each other by the 
medial inserts 40 through which the respective piston shaft(s) 7 extend. 
The piston shafts 7 bear on their axial ends the pistons 34 and 44 
respectively. Instead of using this kind of double cylinders and pistons 
any other suitable arrangements may be applied, for example, those of 
FIGS. 14,15,31 32 or the like. The medial inserts 40 may have an internal 
control chamber 50 if the piston shafts 7 have the inlet flow control 
recesses 45. These control recesses communicate temporarly the inlet port 
104 with a respective one of the working chambers 31 or 41. Air or 
air-fuel mixture under natural or supercharged pressure enters then from 
inlet port 104 over control recess and internal chamber 50 into the 
respective working chamber 31 or 41. An alternative assembly is the 
provision of inlet valves 101 and 102 in the insert 40. These valves may 
by connected by traction spring means 103 and the valve will be closed at 
the respective power strokes. The inlet flow of air or mixture flows then 
from port 104 through the respective opened inlet valve 101 or 102 into 
the respective working chamber 31 or 41. The exhaust ports 39 or 36 will 
be opened respectively when the respective piston 34 or 44 moves close to 
its outer dead point location. The cylinders may be mounted into seats in 
housing 80 and exhaust ports 36 may then lead the exhaust gases into an 
exhaust gas collection chamber 92 in housing 80. This engine requires only 
small space and is very powerful at little weight. Since the double piston 
engines are one cycle engines, it is not required to have cylinders and 
pistons on the bottom portion of the housing because the double pistons 
provide not only thrusting strokes but also tracting strokes to the 
crankshaft 56. It is convenient to set a cooling fan along the axis 86 
because such single fan would then cool the housing as well as all six 
cylinders. The fan can be driven simply by chain, belt or gear from the 
crankshaft 86. Three cylinder two cycle engines were in the fifties in 
Europe called 3=6. This engine of FIG. 20 could then be called 3=12 
because it would have 12 power strokes instead of 3 power strokes of a 
four cycle engine with 3 cylinders. The configuration of this engine 
permit to set the cylinders into the airstream on aircraft and vehicles 
while the housing would remain in the body of the vehicle. 
FIG. 21 is the form of FIG. 9 with the engine of FIG. 10 calculated 
therein, however, in opposite direction. While FIG. 10 starts with 
compression ratio 1, FIG. 21 starts with compression ratio 100 which is 
better for the power stroke. The results of maximally obtainable strokes 
per minute are 1205 in FIG. 21 while they were 929 in FIG. 10. Thus, FIG. 
21 may be more accurate than FIG. 10. 
FIG. 22 shows a Stelzer engine in a 1:1 scale which could obtain the 30 000 
strokes per minute, according to German language Literature. This is 
really a mini engine with very little power. It has Stelzers medial piston 
12 with the pre compression chambers 28,29 about inlet 30. Inlet and exit 
valves 26,27 are shown to operate the outer cylinders 210 and 211 with 
their inlets 6. 
FIG. 23 with cross sectional FIG. 24 illustrate that the engine of FIG. 22 
is not the best solution for the supply of compressed air. The detailed 
calculation in the mentioned DE OS 32 31 718 show that the compression 
piston for the supply of compressed air should have a larger diameter than 
the engine piston 4. Consequently, FIG. 23 shows that the compressed air 
supply arrangement has a compressor piston 33 of a larger diameter than 
the diameter of the engine piston 4. The turbo 68 may be mounted after the 
exhaust to supply pre compressed air either into the inlet of the engine 
or also into the compressor chamber. In order to obtain the high number of 
strokes per unit of time the engine piston and compressor piston must be 
provided with a shaft 38 (or 38 and 37) to extend shaft 38 outwards from 
the cover of the engine to be connected there at 43 with a conrod 46 of a 
crankshaft 63 with a revolving mass 52. Crankshaft 56 may be borne in a 
bearing 35 in crank housing 42. By the provision of the crankshaft with 
the revolving mass the number of strokes of this engine can be multiplied 
compared to the free piston engine without the crankshaft. 
FIG. 25 shows a longitudinal sectional view through a hydrofluid conveying 
combustion angine of the invention. The parts thereof which are already 
known by their referential numbers from other Figures of this 
specification are eliminated from the description of this Figure. FIG. 26 
is the cross sectional view through FIG. 25 taken along the arrowed line 
in FIG. 25 and FIG. 26 should be read together with FIG. 25. The cylinders 
have the inlet valves 26 of FIG. 32. The medial piston shaft 7 is provided 
with stroke cam portions 76,77 to drive with their stroke guide faces 79 
the pistons 24 of the hydraulic pump over the rocker arms 71 the on their 
thrust faces sliding piston shoes 70 while the arms are borne by the cams 
over the rollers 72 on bars 73. The pump pistons 24 are thereby pressed 
into the hydrofluid cylinders 21 an let them be returned to the outgoing 
positions at the opposite strokes of the engine pistons and piston shaft. 
A further specific arrangement of these Figures is that slots 81 are 
provided through the housing or wall of the cylinders to permit the 
application of piston shaft arms 80 provided on the medial piston shaft 7 
and to be extended radially outwardly through the mentioned slots 81. That 
permits the provision of bearing bars on the axial ends of the arms 80 to 
bear pivotably thereon respective conrods 46 or 48 for connection of the 
piston arrangement 7,4.44 with a respective crankshaft which is not shown 
in the Figures. A housing portion 57 may hold the cylinders 2 together. 
The connection of the piston arrangement to the revolving crank shaft 
again serves to make many strokes possible per unit of time and thereby to 
multiply the power of the engine compared to the free piston engines. 
FIGS. 27 and 28 show a modified engine of the invention. Crankshaft 56 
revolves in the crank housing. Crankshaft 56 bears at 63 the conrod 46. As 
a novel arrangement of the invention the crankshaft is subjected to fluid 
pressure pockets from which passages 87 extend through crankshaft portions 
to communicate to a fluid pressure pocket in the eccentric portion 63 of 
the crankshaft. By this arrangement it becomes possible to lead fluid 
fluid under pressure from the outside through a housing portion into the 
crankshaft and bear the crankshaft and the conrods on fields of fluid in 
the mentioned fluid pressure pockets. Another novel arrangement of this 
Figure is, to set a plurality of smaller cylinders as the opposing 
cylinders to the one cylinder 2 with piston 4 therein. The sum of the 
cross sectional areas of the four opposing cylinders with pistons 44 
therein is equal to the cross sectional area of piston 4. Seen are two 
opposing cylinders in the Figures. Instead of two, three or four such 
opposing cylinders may be used, whereby the sum of the cross sectional 
areas of the opposing cylinders and pistons should correspond to the cross 
sectional area of the one single piston 4. The Figures illustrate in 
details how the connection means and locations are provided in order to 
obtain this arrangement. 
As a further novelty of the invention, FIGS. 27 and 28 show dead spaces 
preventing valves 84 and the thereto belonging complementary configuration 
of the top face(s) of the piston(s) 4. Piston 4 has on its top the valleys 
88 of a configuration complementary to the outer diameters of the 
cylindrical bar valves 84. Valves 84 may be revolved or pivoted to open 
and close the the working chamber 1 by the passages 85 through valves 84. 
The radii of valleys 88 correspond to the radii of the outer faces of the 
valves 84. 
FIG. 29 shows in principle the engine of FIG. 26 with the slots 81, arms 80 
and conrods 46. However, FIG. 29 includes a longitudinal sectional view 
through the crankshaft housing with the crankshaft 56 with the revolving 
masses 56 thereon. The Figure further includes a novel valve of the 
invention, namely the inlet valve 26 with a thereto belonging 
complementary configuration of the piston to reduce or eliminate dead 
space. The valve 26 is an inlet valve and is a ball which may be hold by a 
soft spring 89. To prevent dead space the piston head is provided with a 
valley of the form of a hollow ball with a radius which corresponds to the 
radius of the ball of valve 26. A groove 91 may be provided in the piston 
head to take in temporary the spring 89. The piston can now as in FIG. 28 
move so close to the cover of the cylinder that it almost meets the bottom 
of the cover 3 and thereby eliminates or prevents dead space. The 
elimination of dead space in this and in other Figures of the 
specification is desired to obtain high compression ratios and thereby to 
operate the engines with great power and efficiency. FIG. 30 is a cross 
sectional view through the medial plane of the upper portion of FIG. 29. 
Both Figures illustrate that the arms 80 may be assembled to the piston 
shaft 7 by providing a recess through piston shaft 7 and extending arm(s) 
80 therethrough. 
FIG. 30 further shows the important princple that cool fluid inlet ports 
19,20 may be provided in the medial portion of the wall of the cylinder(s) 
in order to lead cooling fluid into the space between the pistons 4,44 and 
around shaft 7, whereby, if the pressure in the cooling fluid is kept high 
enough, back flow of exhaust gases from the exhaust pipe or pipe to the 
turbocharger through outlet ports 6 into the space 5 between the pistons 
4,44 and shaft 7 would be prevented. 
FIG. 31 shows an important modification of FIG. 14 of the invention. 
Instead of providing the piston ring in the piston shaft 7 as it was done 
in FIG. 14, the arrangement of FIG. 31 shows the provision of a radially 
inwardly thrusting piston ring 11 in piston ring groove 10 in the top of 
the cylinder. Piston ring 11 has an inner face 97 to fit and seal on the 
outer face of shaft 7 when shaft 7 meets the inner face 97. The feature of 
this arrangement is that the piston ring, which is not a piston ring any 
more but a seal ring, seals along the entire stroke of piston shaft 7 as 
long as not the control recess 15 moves through the seal ring 11. To make 
an easy assembly of the seal ring 11 possible and to make an accurate 
machining of the seal ring seat possible, the cover 3 may be axially 
divided into two sections as shown by the line therein. After the seal 
ring is inserted into its bed 10 the two portions are set together again. 
A second seal ring 11 may be provided in a second seal ring bed 10 in the 
axially outer portion of cover 3 to seal the outer portion of shaft 7 
against leakage of gas or fluid axially outwards from inlet port and port 
ring groove 9. 
FIG. 32 shows that instead of providing a piston shaft 7 with a control 
recess 15 it is also possible to eliminate the piston shaft 7 and replace 
it by a single concentricly located inlet valve 26 which may be slightly 
loaded by a holding spring arrangement 98.99. The shaft 100 of valve 26 
may be sealed by a seal ring 11 in a bed 10. The tapered seat of valve 26 
and the concentric location of the single inlet valve makes a large cross 
sectional area for the inflow of fluid or air into working chamber 1 (or 
61) possible with an inexpensive and simple valve arrangement. The valve 
26 is opened at the inlet stroke or location of the piston by the suction 
from pressure below inlet flow pressure in chamber 1,61 or by the loader 
pressure in the inlet port(s) 9 and the valve 26 is closed by the higher 
pressure in chamber 1,61 when the compression therein builds up. Thus, the 
valve of FIG. 32 is supposed to open and close automatically under the 
pressures before and behind it alternating with time. Instead of such an 
automatic operation a controlled forced opening and closing could also be 
provided on the axial outer end portion of the shaft 100 of the inlet 
valve 26. 
FIGS. 33 to 36 illustrate a medial insert, for example, as such of FIG. 20, 
in a larger scale and in sectional views. The medial insert 40,140 is in 
these Figures preferred to be divided along line 150 of FIG. 36 in order 
to make the insertion into the cylinder 2,62 possible without dividing the 
pistons 1,61 with medial shaft 7. FIG. 33 would show the medial insert 
portion in a longitudinal sectional view along line 150 of FIG. 36 if the 
insert would not be radially divided allong line 150 into two equal 
symmetric portions. 
FIG. 34 would be a sectional view along the horizontal medial plane of FIG. 
36 or the sectional view through the medial plane of FIG. 33. FIGS. 35 and 
36 are sectional views along the medial arrowed lines of FIGS. 33 or 34. 
FIGS. 34 and 35 show alternatives of valves which may be inserted into the 
medial insert 40 or 140. FIG. 33 shows the longitudinal sectional neutral 
view of the insert 40 without any assembly of valves and needs no further 
description. FIG. 34 illustrates in a larger scale the valves 101 and 102 
and their assessories of FIG. 20. These are already described in principle 
at the description of FIG. 20. FIG. 34, therefore, merely illustrates an 
alternative to the spring means of FIG. 20. Thus, valves 101 and 103 have 
valve shafts with end holders 105. Springs 107 are assembled around 
portions of the shafts of valves 101 and 102. A springs holding housing 
106 surrounds the springs and is provided with outer bords 108 to hold 
thereon the outer ends of the springs 107. The assembly may be done to 
axially passages or inlets 104. In the alternative of FIG. 35 the inlet 
valves 112 are radially arranged. To make their assembly convenient, the 
tapered valve seats 113 seat in valve housing 130 and are able to open and 
close the tapered seat 113. Springs 117 are at one end borne on bords on 
the spring housings 130 and at the other axial ends on the holder portions 
115 which are provided on the shafts 112 of the valves. Stoppers 116 
should be provided, if necessary to prevent an inwards movement of the 
valve heads beyond the internal space 50 of the medial insert 40 which may 
have bords 140 to prevent an inwards movement of the valve housings 130. 
Ignition spaces 109 may be provided in insert 40 and ignition plugs may be 
bolted into the threads 110 of the cylinder's wall 2,62. 
Since the analysis of the engine disclosed that the weights of the 
reciprocating parts should be as low as possible, FIGS. 37 and 38 
illustrate a conrod (pluel, connecting rod) of little weight which can 
also be used in other, for example, in common engines. It is made of FRP, 
for example, of carbon fiber. It has two cylindrical and portions 118 and 
119. A distance bar of a cross sectional configuration of a cross also 
made by carbon fiber, namely portion 120,122 of a cross sectional 
configuration of a cross is inserted between the two cylindrical end 
portions 118 and 119. A holding layer, also of FRP or carbon fiber is then 
led around the periphery of the assembly as shown by referential number 
123. The carbon fiber cloth is glued with epoxy resin or other suitable 
glue and after drying the assembly gives a reliable conrod of a weight 
many times smaller than a conrod of steel. This conrod is also easily 
produced because the carbon fiber will not require expensive machining. 
FIG. 39 shows the calculation table for the calculation of the pressure 
value " P " as an addition to FIG. 10 and it also shows the values 
obtained in the table in a diagram. 
FIG. 40 shows the calculation table for the calculation of the pressure " P 
" and of the pressure " P ". The purpose of this table and of the 
calculation is in details described in my mentioned German DE OS 31 32 
718. FIGS. 41 and 42 show in diagrams the results of calculations by the 
analysis of the engine. FIG. 43 compares in a diagram the values of " P " 
and of " P ". 
FIG. 44 shows in a diagram the increase factor of the power of the engine 
at different pressures of supercharging or pre loading of the ingoing air 
or mixture. 
FIGS. 45 to 50 deal with improvements of free piston engines by the 
invention. FIG. 93 adds a further improvement and will be discussed 
already now at the discussion of FIGS. 45 to 52. Free piston engines, 
which serve as hydrofluid conveying combustion engines, are known for 
example from my U.S. Pat. Nos. 3,260,213 and 3,269,321. Free piston 
engines are also known for example from the West German patent application 
publications 1,451,662 and 3,029,287. The last mentioned publication 
provides a medial piston in a medial chamber between two pistons on yhe 
axial ends. The medial piston and chamber provide the suction and 
pre-compression of fresh air which is then flashed or pressed into the 
working cylinders on the axial locations of the engine. The medial, piston 
has a heavy weight, which has a heavy mass and which prevents high 
frequencies of reciprocation because the heavy mass of the medial piston 
can be stopped at high masses and kinetical energies only with 
difficulties. The pistons of the last mentioned publications thereby tend 
to run against the cylinder heads at high frequencies of reciprocation. 
That limits the frequencies of reciprocation per unit of time and thereby 
limits the power of the engine. The object and aim of the embodiment of 
the invention of the now discussed Figures is, therefore, to overcome the 
difficulty of the known former art and to provide a free piston engine 
with a capability of high frequencies per unit of time. A further aim of 
these Figures is, to improve the hydrofluid conveying combustion engine of 
my former art to a better uniformity of flow and reliability of operation. 
FIGS. 45 and 46 therefore show a cylinder housing 1 which may have an equal 
inner diameter over the entire length in order to make a simple 
inexpensive machining or honing possible. In the medial portion of 
cylinder 1 a control body 15 is mounted, which surrounds a piston shaft 3. 
Piston shaft 3 connects the first end piston 2 with the second end piston 
3. Pistons 2 and 4 as well as the medial control body 15 fit in the inner 
face of cylinder 1, where medial body 15 is fixed, while the pistons 2 and 
4 reciprocate with piston shaft 3 between them in the cylinder housing 1. 
Thereby the cylinders 27 and 29 are formed endwards of the medial portion 
15. Cylinder 29 between 15 and 2, while cylinder 27 is formed between 15 
and 4. These cylinders 27 and 29 alternatingly increase and decrease their 
volumes when the piston 2,3,4 reciprocates in cylinder housing 1. Endwards 
of the straight inner face of cylinder wall 1 there are widened passages 
or annular grooves 29 provided between the sealing face of cylinder wall 1 
and the end covers 8 of cylinder wall 1. Passages 29 are extending from 
the recesses 29 by passages 11 to form the exhaust passages 11. The piston 
shaft 3 is provided with a first control recess 5 and a second control 
recess 6. The medial control body 15 has the bore wherethrough the shaft 3 
fittingly extends and this bore is surrounded medially of member 15 by a 
radially outwardly extending recess 18. The cylinder wall 1 is radially of 
the oitcut or recess 18 provided with an entrance passages 25, wherein a 
one way inlet valve 19 is mounted. Valve 19 opens in the direction towards 
the medial recess 18 and closes into its seat on inlet housing 26 in the 
opposite direction. A spring 20 may hold the valve 19 in closed position 
as long as it is not opened by pressure in the entrance housing 26. A 
stopper arrangement 22,24,21 should be provided to prevent a running of 
valve head of valve 19 against the medial piston shaft 3. The cylinder 1 
and, or medial control body 15 is, are, further provided with either 
ignition means or injection means 16,17 or ignition and injection means 
16,17. These means extend to the respective cylinders 27 or 29 
respectively. It is in accordance with the invention preferred, to mount a 
turbo charger 12 and connect its exhaust gas entrance ports to exhaust 
passages 11. The compressor stage of the turbo is driven by the exhaust 
gases of passages 11 and drive the compressor stage of the turbo which 
takes in fresh air through entrance 13 and delivers slightly compressed 
air, turbo-charged air through charger outlet 14 into the entrance passage 
25 in entrance housing 26 and thereby against the bottom of the one way 
valve 19 to open it, if it gives way. 
The engine of this Figure may operate as follows: At the position of the 
piston 2,3,4 as shown in FIG. 45, compressed air or mixture is present in 
cylinder 29. The fuel is now injected through injector 16, if only air is 
in cylinder 29. But if mixture is compressed in cylinder 29, the 
referential 16 will be an ignition means. In both cases the air and fuel 
in cylinder 29 will now ignite and the gas will burn and expand, whereby 
the piston 2,3,4 is forced leftward. Thereby the piston portion 4 closes 
recess 29 on the right side of the Figure and starts to compress air or 
mixture of air and fuel in cylinder 27. After the expansion stroke towards 
the left end is completed, the left piston portion 2 gives the recess 29 
free to communication with cylinder 29. The expanded and used gas, which 
has giben power at the leftward stroke flows into exhaust passage 11 on 
the left side of FIG. 45 and thereby into the turbine stage of the turbo 
charger 12 to drive therein the compressor stage for the supply of 
prepressure charged air or air-fuel mixture. At the time, when the piston 
2,3,4 was the position as shown in FIG. 45, the medial recess 18 was over 
the first shaft recess 5 in communication with the cylinder 27. Cylinder 
27 was thereby cleaned from burned gases and flashed through with fresh 
air and filled with fresh air. This fresh air or air fuel mixture was 
compressed at the movement of the piston 2,3,4 to the most leftward 
position. There the second annular recess or recess 6 communicated the 
chamber or recess 18 with the left cylinder 29 while the fitting of shaft 
3 in the medial bore in part 15 prevented communication of recess 18 with 
cylinder 27. At this time the recess 18, which receives pre-pressed air or 
mixture from entrance 25 over the then open valve 19, passes the fresh air 
or mixture over the second control passage 6 into cylinder 29 to flush the 
exhaust gas out thereof and to fill the cylinder 29 with fresh air or 
mixture. Thereafter the fuel in air in cylinder 27 which is now highly 
compressed, becomes ignited similar as that in cylinder 29 was ignited 
earlier and the now burning and expanding gases in cylinder 27 now drove 
the piston 2,3,4 rightwards and into the final rightmost position, which 
is shown in FIG. 45. Thereafter the double cycle, which was described, 
starts again. 
During the axial movement of the piston shaft 3 both recesses 5 and 6 move 
at every axial full one stroke through the medial chamber 28 in medial 
control and cylinders separating body 15. That would lead to a backflow of 
a fluid stream from each cylinder one after the other, when the respective 
control recess 5 or 6 would communicate the passage 25 over recess 18 with 
the cylinder 27 or 29 when such cylinder has still a higher pressure than 
is present in entrance passage 25. Such backflow would happen once at 
every stroke of one direction of shaft 3 between piston portions 2 and 4. 
Since such back flow would disturb the effective operation of the engine, 
which the invention clearly discovers, the invention also takes care to 
prevent such a back flow out of cylinder 27 or 29. It does it by the 
insertion of the very important one way valve 19 into the entrance passage 
25 and thereby between entrance passage 25 in housing 26 and the medial 
chamber or recess 18 in the medial body 15. It is convenient to mount the 
one way valve 19 into an entrance housing 26 as shown in the Figure, but 
otherwise it could also be mounted into the medial body 15. For assembly 
of the engine, either at least one of the pistons 2 or 4 is made separable 
and mountable onto shaft 3 or the medial body 15 is divided into two 
halves which are sealing together on each other when they are radilly 
inwardly moved to meet at shaft 3 and then moved radially into the inner 
face of cylinder wall 1 which then holds the two piece medial body 15 
together and under seal. When the entrance housing 26 is inserted as shown 
into an outcut in a portion of the medial body 15, the medial body 15 is 
thereby also fastened axially in place. Inlet valves 9 may be set into 
seats 16 communicated to passages 11 to draw in air at times when there is 
no exhaust gas under pressure in the respective exhaust passage 11. 
The replacement of the pre-compression medial piston portion of the 
publication of the former art by the straight shaft 3 and the medial body 
15 with chamber 18 and the one way inlet valve 19 has very drastically 
reduced the weight of the reciprocating piston and thereby made it 
possible to run the engine with much higher frequencies of reciprocal 
movements then the engine of the mentioned former art could do it. At the 
same time the engine is very much simplified and made inexpensive and easy 
to be produced. A single straight through pipe can now be used as a 
cylinder for the engine and contain both cylinder chambers 27 and 29 as 
well as the medial arrangement 15,18,19. The shaft 3 with piston portions 
2 and 4 can easily be machined and grinded for accurate fit. With these 
improbements the engine has also very considerably reduced its overall 
weight, whereby the aim to use it in aircraft and other vehicles is 
practically fully ontained. Piston set 2,3,4 may on one or both ends be 
provided with an outgoing shaft 7 if so desired. 
FIG. 46 with the thereto belonging cross sectional FIG. 47 improves my 
earlier hydrofluid conveying combustion engine of my mentioned older US 
patents. The engine of these Figures is similar to that of FIG. 46. 
However, the shaft 3 is made longer, the medial assembly 15 is replaces by 
two closing covers 115 and 215 and a cam drive assembly 40,43 is set onto 
the medial portion of shaft 3. Cam plates 40 and 43 are provided with 
radial outer faces 41 or 44 respectively. See hereto also FIG. 47, which 
shows that each cam 40 or 43 consists of a pair of two cams which are 
diametrically and oppositionally directed and located. Instead of 
providing 4 cams as in FIG. 47, it would also be possible to provide thre 
or four cams pairs, which would result in 6 or 8 cams 40,43 etc., The 
cylinder housings around cylinders 27 and 29 are connected to a medial 
engine housing 42, which bears the cylinders 38,138 etc of hydraulic 
pumps. The cylinders may have axes which are normal to the longitudinal 
axis through piston 2,3,4. Pistons 39,139 are able to reciprocate in the 
mentioned pump cylinders 38,138 and they may be provided with pivotable 
piston shoes 37 to be borne on bearing planes of partially plane bodies 
36. Planes 36 may be the ends of arms 32, which are pivotably bearable by 
pins 31 in the walls of the cylinders 27 and 29. They may extend through 
outcuts 33 in medial housing 42. The arms 32 are or may be also provided 
with senser rollers or slides 34, for example with rollers 34 bearable in 
shafts or pins 35. The pressure or pre pressure in the pump cylinders 
38,238,138 and 338--see also FIG. 47--presses the pistons 39, 139, 
239,339--see again also FIG. 47--against the piston shoes 37 and the 
piston shoes 37 into engagement on the plane faces of partially plane 
bodies 36, while thereby the inner ends of the arms 32 are pressed towards 
the cams 40,43 etc. with their sensere 34,35 to run along or be pressed 
against the cam faces 41 or 44. When now the piston 2,3,4 moves leftward 
in FIG. 46, the pump pisyons 39,139 move outwards in cylinders 38,138 and 
thereby towards the shaft 3 because the configuration of the cam faces 41 
of ams 40 permit now this movement because the cam faces 41 now reduce the 
distances from the axis of shaft 3. At the same time, however, the cam 
faces 44 of cam 43 press the pistons 239 and 339 away from the axis of 
piston 2,3,4 and thereby inwardly in and into their pump cylinders 338 and 
238. 
FIG. 48 shows a portion of FIG. 46 seen from the side, whereby the left 
cylinder is seen from the outside and with its cooling ribs 30 and its 
holding levers or pins 31 which pivotable bear the pivot arms 32. 
FIG. 48 shows the further important improvement over my older mentioned 
patents, that it overcomes the ununiformity and partial uneffectiveness of 
my former patents by providing properly configurated cam faces 44 etc. 
The hydrofluid conveying combustion engines of my mentioned earlier US 
patents never obtained their full possible efficiency, because it was 
desired to make the powers of engine piston and pump piston equal, but it 
was never found or disclosed how they could be made equal. The present 
invention now provides the possibility of making them equal by defining 
measures "S", "H", and angle ".theta." of the cam and cam face 43 and 44. 
For the dead space less engine the cam faces 44 shall now correspond 
substantially to the equaitions of FIG. 50. 
FIG. 50 brings the following important equations: 
##EQU24## 
By these equations the cams and cam faces 43 and 44 as well as the other 
cam faces can be made to maintain a power equilibrium between the engine 
piston 2,3,4 and the hydraulic fluid or pneumatic fluid or gas pumping 
pistons 38,138,238 and--or 338 and--or more or some of them. In the 
equations above the following values apply: 
.theta.=local angle of cam face respective to axis of shaft 3; 
.kappa.=polytropic or adiabatic exp nent of gas or air; 
K1=the constant deriving from the design; 
K2=the second constant caoming the from design relation; 
P1=intake or atmosphereic pressure; 
P2=pressure in combustion or compression cylinder. 
H0=zero stroke of piston 2,3,4 
H2=actual stroke of piston 2,3,4 according to FIG. 50. 
When these equations and values are followed, the engine and pump will work 
according to the invention and will thereby work with good efficiency and 
power. 
FIG. 49 shows the arms 32--they are double arms laterally of plane face 
body 36 and holder 45 on the cylinder--in a view along the arrow above 
FIG. 50 from above. There is nothing special in this Figure, but it shows 
in the view, what FIG. 50 could not show on the plane sheet of the paper. 
FIG. 51 corresponds in principle to FIG. 15. However, FIG. 51 shows the 
improvement or alternative that a cooling fluid supply chamber 19 is 
provided which blows cooling fluid through ports 6 into medial chamber 59 
between pistons 4 and 64. This cool fluid is then partially led through 
passge 160 into the hollow piston shafts 60 and through shafts 7 and 
partially out of space 59 through exit ports 20. It is also possible to 
send all cooling fluid through the shafts 60,7 or all cooling fluid out of 
chamber 59 through outlets 20. 
FIG. 52 is a longitudinal sectional view through a double piston engine 
with variable pressure ratio. Instead of providing this embodiment of the 
invention to double piston engines it could also be applied to single 
piston engines with a single piston in the respective cylinder 2 or 62. 
The application to a single cylinder is, however, not shown in this Figure 
because it is easily understood to do so by eliminating the bottom portion 
or the top portion of the Figure. The principle of this embodiment of the 
invention is, thereby, applicable also to common combustion engines or 
engines, devices, pumps or motors of the known art with a reciprocating 
piston in a given cylinder. 
Crankshaft housing 57 is provided with a cylinder guide or cylinder guides 
160 wherein the respective cylinder 2,62 is axially moveable along its 
longitudinal axis. A compression ratio adjustment housing 161 is provided 
to the engine. It forms a space or slot 161 wherein the crankshaft housing 
57 is provided. Adjustment controllers can be provided to the adjustment 
housing portions 161 to move them axially towards each other or away from 
each other in the direction of the arrows in the Figure. The portions 161 
keep axially the respective cylinder 2 or 62. By moving the adjustment 
holders 161 axially, the respective cylinders 2 and/or 62 are also moved 
axially along the arrows in the Figure. Thereby the compression ratio is 
varied because the distance between the pistons 4,64 and the covers 3,63 
varies, whereby the compression ratio is varied in accordancee with the 
definitions of FIGS. 3 and 5. The other parts of the engine of this Figure 
are known from FIG. 17. 
FIG. 53 corresponds in principle to FIG. 46. However, instead of a 
hydraulic piston a gas pressure supplying piston 4 with shaft 164 is 
provided in the fluid flow creating cylinder 21. Pluralities of such 
pistons and cylinders are commonly provided and two of them, opposingly 
directed, are shown in the Figure. Inlet valves 84 are provided and the 
pistons have respective configurations of the head faces as described at 
hand of FIG. 28. Passages 165 are provided to prevent varying pressures 
below in cylinders 21. The piston shoes 70 are inserted into shafts 164 
instead of into hydraulic fluid pressure pistons as in the other 
respective Figures. The embodiment of FIG. 53 of the invention is very 
convenient as a compressed air providing engine. It is of little weight 
and inexpensive in production. The diameters of the cylinders 21 and of 
the pistons therein make different ranges of air pressure possible. 
FIG. 54 corresponds in principle to FIG. 53. However, piston shafts 164 
retracting guide rails 170 with guide faces 171 which guide the bars 73 of 
the rollers 72 in the retraction stroke. Thereby the pistons in cylinders 
21 are forced inwardly and obtain the ability to suck fluid into the 
working chambers in the fluid flows producing cylinders 21. Oppositionally 
acting guide rails 172 may be provided on engine shaft 7 if they are 
angularly spaced from the guide rails 170. 
FIGS. 55 and 56 show the engine of FIG. 15 in 6 locations of the piston in 
longitudinal sectional views. Therein arrows are provided which with teir 
thickness and length indicate the concentration of pressure in the 
respective working chamber. Below the sectional views diagrams are 
provided which show the expansion pressures, the compression pressures, 
the medial velocity of the piston, the braking velocities and the points 
"G" where the braking of the running piston starts. More details thereof 
are, again, found in my mentioned German DE OS 31 32 718. It should be 
noted, that according to this invention the pistons should be provided 
with connection means to make the connection of a conrod to a revolving 
crankshaft with a revolving mass possible in order to make an increase of 
the number of strokes per unit of time possible. That is indicated by the 
insertion of connecting portion 343 into or onto the end of one of the 
piston shafts. A cross pin 43 may connect the connecting portion 243 to 
the respective conrod. 
FIG. 57 shows in a longitudinal sectional view the arrangement of means to 
prevent backflow of hot gases from the exhaust or fluid line to the 
turbocharger into the interior space between the pistons 4 and 64. For 
that purpose the one way check valves 306 which may be loaded by springss 
406 are set into the exhaust passages 6 and 66 or one of them or into a 
combined exhaust passage 666. These valves prevent that exhaust gas which 
has already left the respective cylinder 1 or 61 after the end of the 
exhaus stroke could flow back from the collection chamber 19-16 or 319-316 
into the space 59 between pistons 4 and 64. The arrangement of this valve 
is important to prevent excessive heating of the walls 2,62 of the 
cylinders and of the pistons 4,64 or the medial piston connecter portion 
60. 
FIG. 58 shows in a longitudinal sectional view another provision to prevent 
excessive heating and back flow in and into the chamber 59 between the 
pistons 4 and 64. In this Figure the medial piston rod or connecting 
portion 60 is replaced by a medial portion 464 of larger diameter or by 
two medial portions 404 and 464 of a larger diameter. The mentionaed 
larger diameter is so large that the outer diameter is so big that only a 
narrow space 59 remains between the medial portions 404,464 or one of them 
and the inner diameter of the cylinder 2,62. Thereby it is secured that 
only a small amount of fluid can flow back from the exhaust or from the 
collection chamber 19 in collecter 16 into the space 59 between the 
pistons, the medial portion(s) and the wall(2) of the cylinder. That 
prevents uniniformity of exhaust flow due to fluctuating flows into space 
59 and in addition it permits the application of a larger cooling surface 
from the interior space of the medial portion(s) 60,404,464. For 
convenience of manufacturing the circular portions 404 and 464 may be of 
different diameters to permit the one of them to fit into the other. A 
holding means, a rivet 411 in the Figure, may be set to hold both medial 
portions 404 and 464 and thereby the pistons 4 and 64 together. 
In FIGS. 57 and 58 it is of further interest that the cover 3,63 should 
have an annular recess 315 communicated to inlet 309 in order to permit a 
large cross sectional area for the inflow of the fluid when control recess 
15 meets the annular groove 315. The recess 15 should also be an annular 
groove and the faces of the recess 15 should be taperedly inclined in 
order to abtain a streamlined flow to prevent losses by friction and by 
directional changes in flow. To prevent break of piston rings forward 
extensions of shaft or cover 7 or 3 and/or extensions (in axial 
directions) of the grooves or recesses 15,315 may be applied to obtain a 
gradual application of seal and deformation of the piston for sealing 
purposes. These arrangements should also be done in FIGS. 14,15 and the 
respective other Figures; the referentials 315 and 15 will indicate these 
applications in the mentioned other Figures of the specification. Also 
applied in FIG. 58 and in the respective other Figures, like Fig. 15 etc., 
are the cylindrical face portions 262 of the portion of the wall(s) 2,62 
of the cylinders 1,61 between the exhausts 6 and 66. Thise face portions 
262 on the medial wall portions 362 have the purpose to guide the 
respective piston 4 or 64 at the respective portion of its (their) stroke 
(strokes). 
FIG. 59 is a longitudinal sectional view through a portion of another 
hydrofluid conveying combustion engine of the invention. It is related and 
partially similar to FIGS. 45 and 46 to 48; however, the cams on the 
medial piston shaft 7 are different and serve different purposes. The cams 
576 on the medial shaft 7 between the pistons 4 and 64 have in this 
embodiment pump piston stroke guide faces 531 of a very different 
configuration for a very different purpose. The cams form portions and 
guide faces 530 with a steap angle at the begin of the expansion stroke of 
the engine piston 4,64 and steap rear portions 532 near the ends of the 
mentioned expansion strokes while in the middle between portions 530 and 
532 the flatter portions 531 with less steep inclinations are provided. 
This arrangement serves to obtain equal rate or almost equal rate of flow 
in the hydrofluid pump cylinders 21 over the entire length of a single 
power stroke of an engine piston 4 or 64. The Figure shows only those cams 
and stroke guide faces which are visible in the section, while those 
angularly spaced thereto for the reverse direction of the engine strokes 
are indicated only by referential 577. It is known from FIG. 47 that these 
may be 90 degrees angularly spaced relative to cams 576 of FIG. 59. 
FIG. 60 is a diagram and explains the values of the cam arrangement(s) of 
FIG. 59. The diagram of FIG. 60 has as the x-axis the stroke "H" of the 
piston 4 or 64 of the engine portion of FIG. 59. The velocity Vpcon of the 
engine piston is shown thereover in the direction of the y-axis. Please 
note, that FIG. 59 shows that the piston 4,64,7 is connected to the conrod 
55 of a crankshaft and that the crankshaft revolves with a given RPM 
whereby the velocity of the piston at any location of its stroke is 
defined and calculable from the rotary angle alpha of the crank of the 
crankshaft. A straight face, inclined relative to the axis of the piston 
would bring the dotted lines of pump stroke Spp of the pump pistons in 
cylinders 21. Such strokes would give a straight face on the cam(s) but it 
would bring a very ununiform flow in the cylinders 21 whereby all piping 
or hosing connections on cylinders 21 would break. The present invention 
discovers this important occurrance and takes the consequences thereof 
thereby that the cam's stroke face gets the mentioned portions 530,531 and 
532 also shown in FIG. 60. 
FIG. 60 shows by a dotted line also the medial velocity Vm of the piston(s) 
of the engine. The actual velocity Vpcon is very different therefrom. It 
is slower at the beginning, higher at the medial portion and again slower 
near the end of the expansion stroke. To nivelize this matter to a uniform 
medial piston speed in the pump pistons 24, the cam's piston stroke guide 
faces must get the steep portion 530 to complement the slower Vpcon and 
get the steeper portion 532 to complement the slower speed portion of 
Vpcon close to the end of the piston stroke of the engine with the flatter 
medial portion 531 therebetween. The Figure shows a stroke of the engine 
piston 4,64 of 54 mm. The crankshaft is calculated to have the conrods 
centered on a radius of 27 mm around the concentric axis of the crankshaft 
and the length of the conrod=distance between the center axes of the eyes 
of the conrod=is calculated to be 110 mm. This corresponds to one of the 
Yamaha motor bike engines. The guide face "Spp"=530,531,532 would then 
bring also 54 mm stroke to the pump pistons 24 and the velocity of the 
pump pistons 24 would then be equal at the entire stroke to the medial 
velocity Vm of the engine's piston(s) instead to the actual velocity Vpcon 
of the engine's piston(s) 4,64,7. This is accurate, if the pump pistons 
244 meet the stroke guide face 530-532 in points or parallel lines as 
shown in FIG. 60. For actually applied rollers 72 respective adjustments 
might be required. Since commonly the pressure in the pump cylinders 21 is 
higher than the pressure in the engine's cylinders 1,61, a shorter stroke 
of the pump pistons 24 is suitable. FIG. 60 shows therefore, a second 
curve "Spp" for a stroke of 13.5 mm which means for a four times shorter 
stroke. In summary, the courves Spp are actual sizes relative to the 
written dimensions in the Figure, for the actual machining of the piston 
stroke guide faces of cams 576 of FIG. 59. The other parts of FIG. 59 
correspond to respective part of others of the Figures of the 
specification. 
FIG. 61 shows in a longitudianl sectional view a modification of the cam 
arrangement to a high pressure hydrofluid conveying hydrofluid conveying 
combustion engine. The earlier Figures have rollers 72 which meet the 
piston stroke guide faces of the cams only in a line contact. Line contact 
has only a limited bearing capacity. To obtain a higher pressure in the 
hydraulic pumps the line contact should be changed to a face contact which 
permits a higher bearing capacity. To obtain that in FIG. 61 the pistons 
24 bear therein pivotable piston shoes 321 with plane slide faces which 
are complementary configured relative to the piston stroke guide faces 331 
of cams 376 whereon they actually slide. The piston stroke guide faces 331 
are, consequently, also plane faces whereby the stroke cams 376 form 
inclined plane faces which are angularly inclined relative to the 
longitudinal axis of the piston(s) of the combustion engine. The Figure 
also indicates by 377 the cams for the oppositionally directed stroke. 
Shown are also the hydrofluid cylinder spaces 721 in cylinders 21 with the 
outlets or inlets 721. The arrangement of the Figure has the feature that 
it can operate the pumps with higher pressure because of the higher 
bearing capacity of the faces bearing instead of the lines bearing. 
However, it has the disadvantage that the delivery of fluid out from the 
pumps 21-24 is very ununiform because of the straight plane inclined faces 
331 of these cams 376 of this Figure. 
FIG. 62 with the thereto belonging cross sectional FIG. 63 through the 
arrowed line of FIG. 63 partially overcome the problem of the ununiformity 
of flow of FIG. 61. FIG. 62 is a sectional view through FIG. 63 along the 
arrowed line B--B in FIG. 63. The pistons 24 have again, as in FIG. 61, 
piston shoes with slide faces which are complementary configurated 
relative to the respective piston stroke guide faces. Thus, also this 
arrangement is capable of high pressures because faces slide on faces 
instead of lines rolling on faces. The difference, however, compared to 
FIG. 61, is that the stroke guide faces 481 to 488 are configurated as 
portions of faces of cylinders and that the thereto complementary 
configurated slide faces 490,491 of the piston shoes 321 are portions of 
outer faces of cylinders or of round bars. The stroke faces are provided 
on the cams 476,576,676 and 776. The stroke face 481 is formed with radius 
E around axis A; stroke face 485 is formed with radius F around axis B; 
stroke face 482 is formed with radius G around axis C; stroke guide face 
487 is formed with radius H around axis D; stroke guide face 488 is formed 
with radius N around axis J; stroke guide face 484 is formed with radius O 
around axis K; stroke guide face 486 is formed with radius P around axis L 
and stroke guide face 483 is formed with radius Q around axis M. In 
actuality the radii are shorter than shown in the Figure by which the axes 
are more close to the shaft 7 of the engine. The Figure shows four concave 
piston stroke guide faces and four convex piston stroke guide faces. One 
convex face forms together with a concave piston stroke face a piston 
stroke faces pair. The next speciality of these Figures is, that the 
pistons of the respective stroke face pair form together a single pump in 
which both pistons pump into a common pumping chamber. For example, 
pistons 24 and 324 are one piston pair and pistons 724 and 824 are another 
piston pair of the respective pump of a piston pair. Each pump has thereby 
two pistons for a common pumping chamber with one of the pistons sliding 
on a concave piston stroke guide face and the other piston of the same 
pair sliding on a concave piston stroke guide face. The piston shoes have, 
consequently, per each pump chamber with two pistons of the respective 
piston pair a concave slide face and the other piston shoe a convex slide 
face, either 490 or 491 to be complementary configurated relative to the 
respective piston stroke guide face whereon the respective piston shoe 
slides. 
The important feature of this embodiment of the invention is that a concave 
cam face and a convex cam face act together into a single common pump 
chamber. The common pump chambers per piston pair are shown by 492 and 493 
in FIG. 63. One of the convex or concave piston stroke guide faces thereby 
has a relative steep angle of inclination at the start of the stroke and 
the other at the end of the stroke of the engine, while in the middle area 
of the stroke both faces are relatively little inclined relative to the 
axis of the engine's piston shaft 7. Since both pistons of the pair act 
together into the same chamber the sum of the delivery of both pistons of 
the same pair is more uniform than that of FIG. 61 and nears the 
uniformity of flow of FIG. 59 with diagram 60. A full uniformity is, 
however, not easily obtainable with two pistons in a singe pumping 
chamber, but is almost perfectly obtainable by a plurality of more than 
two pistons per common pumping chamber. Chamber portions 492 are 
communicated to form a common chamber by passage 802. Each common pumping 
chamber has at least one inlet valve 803 and one outle valve 803. Each 
common chamber has an inlet passage 804 and an outlet or delivery passage 
801 or 805. 
FIG. 64 shows in a longitudinal sectional view that it is preferred to set 
a turbocharger between the exhaust port and the inlet ports in FIGS. 
15,14,17,20,57,58 and the other respective Figures. Exhaust port 19 
delivers the exhaust gases into the entrance 441 of the turbine of the 
turbo charger 440. The pre compressed air or air-fuel mixture leaves the 
compressor stage of the turbo 440 to flow over the pipes or fluid lines 
442,443 and their ports 444,445 into the entrance ports 9 of cylinder 
chambers 1 and 61 of the engine of the invention. 
The embodiment of FIG. 65 shows a crankshaft arrangement of the invention. 
The aim of this arrangement is to provide a crankshaft which is easy in 
production without Jigs or machines for eccentric machining. At the same 
time it may or shall have means to run at least one or a plurality of 
pumps. The housing 501 carries in bearings 502 the revolvable shaft 503 
with axis 521. The ends of the shaft 503 hold the crank portions 514. Key 
means 511 and holders 510 may be provided if so desired to fasten the 
crank portions 514 to the shaft 503. Actually the crank portions may be 
fastened by a press fit by warming the crank portions for assembly or by 
cooling the shaft for the assembly. The key and holder can then be spared. 
The crank(s) 514 are now simple forged or casted parts which can be 
drilled or bored by a boring machine with parallel axes of the bores. One 
bore for the fastening on the shaft and the other bore for holding a 
conrod bearing bar 506 therein. Holding means, for example, rivets 509 may 
be provided if so desired. The crank portion 514 has thereby a medial 
portion which is borne on the shaft 3, one radial portion 505 which bears 
the conrod bearing holder 506 with axis 522 which is distanced radially 
from axis 521 of shaft 503 but parallel thereto and in the diametrically 
opposite direction the mass or counter weight portion 504. The crank 514 
on the upper portion of the Figure is shown to be 90 degrees turned 
relative to crank 514 off the bottom portion of the Figure. The 90 degrees 
turning is, however, only done by way of example. The cranks could also be 
equally angularly set or spaced angularly under a different angle, for 
example, 180 degrees or any other suitable degree. The simplicity of the 
design makes it possible to assemble onto the simple straight shaft 503 
any desired drive means. In the Figure a medial gear 512 to drive 
accessories is assembled and endwards thereof are symmetrically eccentric 
cams 515 to 518 assembled to have outer faces to form piston stroke guide 
faces 519 and 520. Faces 519 form one stroke pair and faces 520 form 
another stroke pair. Each stroke pair might also consist of one stroke 
face 519 and one stroke face 520. These stroke faces may serve to guide 
pistons or piston shoes of a hydraulic or pneumatic pump arrangement which 
shall be driven by the conrods 507 which connect to the pistons of a 
respective combustion engine. This crank shaft of FIG. 65 is especially 
suitable and inexpensive to be assembled to the engines of FIGS. 15,64 and 
others of this specification. The crankshaft of this Figure can be 
machined on simple machine tools in small workshops. 
In FIGS. 15,51 and the thereto related Figures, the pistons 4 and 64 might 
be combined to a single axially very short piston, just long enough to 
open and close the combined exhaust of FIG. 57. The stroke of the piston 
would then have to be substantially doubled if the cylinders remain of 
equal lengths. This arrangement would still further reduce the weight of 
the piston, but it is not shown in the Figures. In FIG. 29, bottom portion 
of the Figure, it is shown how the connecting rods 47 to 48 may be set 
directly onto a single eccentric portion of a crank shaft. Plural conrods 
46,47 or 46 to 48 or more may by this way combine the piston strokes of 
multiple double or single piston engines, like, for example, that of FIG. 
20 to working actions one after the other in timed relation relative to 
each other. The Figure shows how such arrangement may be obtained in a 
simple and inexpensive device and design. 
In FIG. 58 it is of value that the outer diameter(s) of the medial 
connecter(s) 404,464 is (are) smaller than the diameter(s) of the seal 
portions of pistons 4 and 64. Because otherwise the required narrow space 
459 would not appear between the pistons 4 and 64. Without such narrow 
space the entire length of the outer face of connecting portion 464 would 
run along the cylinder and wear there, by which it would run through a hot 
portion of the wall of the cylinder(s) 2,62 and might weld there under 
heat expansion or contraction under periodically varying heats. 
The embodiments of this specification show samples of actuall design or of 
prospected designs. The embodiments should be evaluated in combination 
with the analysis of the engine of this specification. Many portions of 
the analysis are entirely exact. Others are present attempts to advance 
towards a better knowledge of the acting medial pressures " P " and " P ". 
The attempts to advance to a better knowledge can not presently be final 
and exact solutions. They may become improved with time in the future. 
Thus, only those portions of the analysis which are assumed to be exact 
should be used in exact values while those portions of the analysis which 
are only present attempts to advance towards a better knowledge should not 
be considered to be final exact values or solutions. 
The pump portions may be hydraulic fluid pumps or pneumatic pumps or 
compressors respectively. They may also act as pneumatic or hydraulic 
motors to drive or to start the portions of the combustion engine. The 
appended claims should be considered to be portions of the description of 
the preferred embodiments of the invention and/or portions of the summary 
of the invention. 
FIGS. 66 and 67 show longitudinal sectional arrangements through an ultra 
power engine of the invention for which international priority of German 
(FRG) application P 36 20 691.1 of Jun. 20, 1986 is claimed. In FIG. 67 
two double acting pistons, running in respective cylinders, are combined 
by a common crankshaft for operation in unison. FIG. 66 shows the 
arrangement in one of the cylinders in a larger scale than the 
arrangements are illustrated in FIG. 67. This engine is called 
"ultra-engine" because it produces greatest power in a lowest weight 
device which can be easily and inexpensively produced. The aim of this 
engine is to produce an engine for the twentieth of the costs of the 
Tornada accessory shaft gasturbines of the Tornado fighter plane of 
Europe. 
Both Figures will now be described together, since FIG. 67 has two devices 
of FIG. 66 with equal referential numerals. In FIG. 67, however, some of 
the portions have equal end numbers in the end digits but pre digits 1,10 
or the like to explain different temporary locations. The equal end digits 
define that the parts are equal to those in the other cylinder of FIG. 7 
or of FIG. 66. 
In the respective cylinder 2,62 the double piston reciprocates and has the 
respective piston shaft 7,107 with one piston 4,104 on one of its ends and 
another piston 64,164 on the other of its ends. The pistons are fitted in 
the cylinders and seal on the inner faces of the cylinders, while piston 
rings may be inserted into the respective pistons to seal the pistons 
along the walls of the cylinders. The speciality of these Figures is that 
according to these embodiments of the invention the piston shaft 7,107 has 
a medial flow control recess 15,115 which extends through the outer face 
of the piston shaft into the piston shaft and which is located 
substantially axially in the middle of the piston shaft and thereby 
axially seen also in the middle between the pistons on the ends of the 
shafts. 
A medial housing 40, is flanked axially by intermediate bodies 3,63 and 
axially endwards of these intermediate bodies are the cylinders 
2,62,102,162 provided. Between the medial housing 40 and the intermediate 
bodies 3,63,103,163 are seal ring beds 53 provided which contain the seal 
rings 54 and 55, respectively. These seal rings are provided with inner 
faces which seal along the outer face of the respective piston shaft 
7,107. The mentioned seal rings have an inner stress which spans them 
radially inwardly for close engagement and seal on the outer face 66 of 
the respective piston shaft. This spanning force may be assisted by 
pressure in fluid in a respective neighboring cylinder chamber by means of 
a respective passage 41 which leads the pressure from the respective 
cylinder chamber into the seal ring bed and onto the radial outside of the 
seal rings. The medial housing 40 is provided with an entrance passage 9 
which forms a chamber portion radially around the piston shaft 7 or 107. 
The arrangement may be held together by the fastening or holding means, 
bolts, nuts, flanges etc., 20 and 21. Holding means 10 may be provided on 
the medial housing, cylinder or intermediate body for the insertion of 
ignition or fluid injection means, for example, 11. The piston rings 52 
are provided in piston ring beds 51. The outer diameter of piston shaft 
7,107 is 88. The medial recess 15,115 ends in control corners 81 and 82. 
The pistons may be fastened to the piston shaft by holding means 
16,17,12,14,14 or the like. The cylinders have exhaust ports 6, In FIG. 66 
the piston arrangement is in the upwardmost location, at which the exhaust 
ports 6 in cylinder 2 are opened because the piston 4 run upwards over 
them. The axial length of the exhaust ports is defined by 67. The length 
of the piston stroke is 67 plus 63 with 63 beeing the length at which the 
respective working chamber 1,61,101,161 is closed during the piston 
stroke. By 61 the distance between the piston and the respective corner 81 
or 82 of the control recess 15,115 is defined. The control recess has an 
outer diameter 89 which is considerably smaller than the diameter 88 of 
the piston shaft. 
The length and location of the control recess is such, that the recess 
15,115 opens a communication between the entrance port 9 and the 
respective cylinder chamber 1,61,101,161 near the ends of the piston 
strokes. Thus, in FIG. 66 entrance port 9 is communicated by control 
recess 15 to the cylinder chamber 1. Fluid enters at this location and 
time from entrance port 9 through control recess 15 into cylinder chamber 
1 and at the same time the old fluid of cylinder 1 is exhausted through 
exit port 6. It is preferred to lead fresh fluid under a certain loading 
pressure into entrance port 9, for example, by a turbo charger. When the 
piston assembly starts to move down in FIG. 66 and the piston 4 runs over 
the exhaust port 6 to meet the cylinder's wall at 61, the exhaust port 6 
is closed and at substantially the same time the control corner 81 of the 
control recess 15 meets the respective inner face of the respective seal 
ring 54,55 to close and seal the entrance port 9 from the chamber 1. 
Similar actions take place at the bottom near location of the piston 
assembly with the then with entrance port 9 communicating and 
discommunicating cylinder chamber 61 on the other side of the medial 
housing 40. In FIG. 66 the engine is ready for ignition or fuel injection 
which will then lead to the expansion of the charge under pressure for 
driving the piston 64 downwards in the power stroke. 
In FIG. 67 two assemblies of FIG. 66 are assembled side by side. Connecting 
rods (conrods) 14,114 connect the respective piston assemblies to the 
common crankshaft 19. The eccentric bearing portions 26 and 126, which 
bear the outer ends of the connecting rods 14,114, are angularly turned 
ninety degrees relatively to each other when seen along the axis of the 
crank shaft 19. The crank shaft is revolvably borne in bearings 25 in 
crank housing 8. There may be two crank shaft portions connected angularly 
together by connecting means 28 and the crank shaft has counter weight 
masses 27,127 relative to the eccentric bearing portions 26 and 126. This 
arrangement secures a certain timed running relation of the piston 
assembly strokes in the two cylinders of this engine. A turbo charger, not 
shown in the Figure, is connected with the delivery line to entrance 30 of 
the entrance ports 9 which are thereby combined to a common entrance 9 and 
a common loader or turbo before entrance 30. Exhaust collection chambers 
23 take in the exhausts from the exhaust ports 6 and transfer the exhausts 
to the turbine of the turbo charger before entrance 30. Cooling fluid 
chambers 24 in cooling housings 29 or respective cooling ribs for air flow 
cooling may be provided on the cylinders. In the arrangement of FIG. 67 
the right portion shows the arrangement of FIG. 66 with the piston 
assembly at this moment of time located as described at hand of FIG. 66. 
Since the eccentric bearings of the crank shaft are 90 degrees turned 
relative to each other, the engine of FIG. 67 has four power strokes per 
each revolution. These power strokes act with 90 degrees turn of the crank 
shaft one after the other. Accordingly one sees in FIG. 67 the cylinder 
chamber 1 at exhaust and fresh loading timing, the chamber 61 ready for 
fuel injection or ignition, cylinder chamber 101 under compression of the 
gas in it and cylinder chamber 161 in the timing of power stroke. The 
arrowed lines in the Figure show the movements of flow of gas or fluid. 
Thereby 31 indicates the flashing of the cylinder chamber by fresh fluid 
from entrance 30,9 in combination with the exhaust 32. The compression of 
the fluid or gas is indicated by 33 and the power stroke of the charge is 
indicated by referential numeral 34. 
With exclusively means of low weight, compact design and four power strokes 
per every single revolution, this ultra power engine obtains a superiorly 
high power per weight and size of the engine unit. 
The engine of FIG. 66 thereby is: 
a double piston device with endwards of a medial housing provided cylinders 
with a therein reciprocating piston assembly, consisting of pistons on the 
ends of a piston shaft between said pistons, exit ports on the axial outer 
end portions of the cylinders, inlet passage means in the medial housing 
with control means for the inflow of fluid into the respective cylinder 
and an improvement, 
wherein the improvement comprises, in combination, 
a control recess provided substantially in the middle between the pistons 
and on the piston shaft between the pistons extending radially into the 
piston shaft and having a length 66 while the piston assembly has a stroke 
of the length 63 plus 67 with the recess 15,115 ended by control corners 
81,82, and a portion of the medial housing surrounding portions of the 
piston shaft and sealing along the respective portion of the outer face 66 
of the piston shaft, 
whereby at the outer ends of the piston strokes the entrance port 9 of the 
medial housing communicates alternatingly with one of the cylinders while 
at the strokes between the end portions of the strokes the cylinders are 
discommunicated from the entrance port in the medial housing. 
FIG. 67 defines 
a double acting device as in FIG. 66, wherein a plurality of piston 
assemblies are provided in a plurality of cylinder arrangements of FIG. 
66, one end of each piston assembly is connected by a connecting rod to an 
eccentric portion of a crank shaft and the eccentric portions of the 
crankshaft are angularly spaced by a number of degrees suitable to the 
number of piston and cylinder assemblies in order to let the piston 
assemblies act one after the other in timed relation at a single 
revolution of the crank shaft. 
FIG. 67 also defines 
a plurality of double acting piston assemblies in a respective plurality of 
cylinder and medial housing arrangements with the piston assemblies 
connected to a common crank shaft to move the piston assemblies per each 
revolution of the crank shaft in timed relation one after the other and 
wherein exhaust chambers collect the exhaust gases from the exit ports 6 
of the cylinders to lead the exhaust to a turbine of a trubo charger while 
the compressor of the turbo charger is communicated to the entrance port 9 
of the medial housing to press fluid under pressure through the medial 
housing over the control recesses in the piston shafts into the respective 
cylinder of the device in timed relation into one of the cylinders after 
the other. 
In FIG. 68 the engine has a crank shaft 503 with counter weigth 504 and 
connecting rod 507 in crank housing 501. Connecting rod 507 connects to 
piston shaft 607 by connecters 647,648,747 and 748. Piston 607 has the 
piston shaft 607 and the rear piston 664 while the front piston 604 is 
mounted on the front of piston shaft 607. The medial housing 640 and 
inserts 641,642 between the medial housing 640 and the cylinders 602,662 
surround the piston shaft 607. Seal beds 643 are provided between the 
medial housing and the inserts while seal rings 644 are inserted into the 
seal beds. The seal rings have inner faces which slide and seal along the 
outer face of the piston shaft 607. The pistons 604,664 reciprocate in 
cylinders 602,662 and seal on their inner faces while piston rings may be 
inserted to improve the sealing. The cylinders are provided with exhaust 
ports 6,66 similar as in others of the Figures and exhaust collection 
chambers 619 in exhaust housings 616 collect the exhaust from the exhaust 
ports and lead it over passages 442,443 to the turbine of the turbo 
charger 440 to drive the turbine while the compressor of the turbo 440 
presses gas or air out of its delivery port 654 into the entrance chamber 
653 of the engine. Inlet valves 650,651 are provided between the entrance 
chamber 653 and the cylinder chambers 601 and 661, respectively, while 
springs 652 are set to close the inlet valves 650,651. If the pressure in 
the cylinder chambers becomes smaller than the pressure in the entrance 
chamber 653 the inlet valve opens to the respective cylinder chamber with 
the lower pressure. Holding means, for example, threads 645,646 are 
provided for the insertion of injection or ignition means. Since piston 
664 may be intergral with piston shaft 607, the piston shaft can be easily 
inserted into the sealing and fitting bores in the medial housing and in 
the inserts. The other piston 604 can then be srewed or held by a nut 649 
on the other end of the piston shaft. It is possible to make the shaft 607 
hollow and to insert the holder 647. This engine works similar as that of 
FIG. 66, however, the flow control recess of FIG. 66 is here in FIG. 48 
replaced by the multiple inlet valves 650 and 651. While for each cylinder 
only one inlet valve and one holder thread is shown in the Figure, a 
plurality may actualy be applied angularly spaced around the axis of the 
engine. 
In FIG. 68 as well as in FIGS. 66,67 and others, it is preferred to obey 
the rules of FIG. 13 and of its explanations in order to make the counter 
weights of the crank shafts as small as possible in order to obtain the 
high power output of the ultra engine by a small weight and size of the 
engine assembly. 
Since the invention is still more in detail described in the appended 
claims, the claims should be con sidered to be also a portion of the 
description of the invention and its preferred embodiments.