DE1957174A1 - Method and device for converting thermal energy into mechanical energy - Google Patents

Description

DR. MÜLLER - BORE DIPL.-ING. GRALFS DIPL.-PHYS. DR. MANITZ DIPL.-CHEM. DR. DEUFEL

PATENT LAWYERS

Hünchen, the 1 3rd NGV. 1969

Hl / th - Δ 2050

Atomic Anergy of Canada Limited Ottawa, Ontario, Canada

Method and device for converting thermal energy into mechanical energy

The invention relates to a method and a device for converting thermal energy into mechanical energy. The invention is applied in particular, but not exclusively, to the use of heat in the case of a not relatively high temperature.

The subject of the application arose from the need to create a small heat engine, the thermal one Energy resulting from the decay of a radioisotope source

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is derived, converted into mechanical and electrical energy. Such a requirement was in anticipation suggested that certain navigational aids, i. H. Lamps, fog detectors, fog signals, radio beacons and the like, can be left unattended for long periods in remote locations.

It was designed to achieve operational reliability, safety and maximum use of a relatively expensive radioisotope source, Considered the minimum conversion efficiency at 20%, the output power for one embodiment of the order of 600 W and the maximum temperature of the available thermal energy approximately at 600 ° 0 to be set.

With these goals in mind, all available methods for energy conversion were examined in detail. This is described below and the invention at the end as a particularly advantageous method of energy conversion specified. It should be noted, however, that the invention is generally applicable and not limited to applications is limited as given above.

It is an object of the invention to provide a method for converting thermal energy into mechanical energy in a continuous, closed-loop process using a working medium which is gaseous at any time.

In accordance with this aim, the method comprises the steps of: compressing a gaseous medium in a first space with a lower mean temperature and a lower mean pressure in a compression ratio "r ^w , heat at substantially constant pressure dem

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to supply compressed medium, further heat from a Source the compressed medium "at essentially constant-T-em To add volume to the medium in a second room with a high average temperature and a high average Relaxing pressure in an expansion ratio "E", where "E" is greater than "r", an amount essentially of heat equal to the first heat supplied to withdraw the expanded gas at essentially constant pressure, and the medium to cool at substantially constant pressure, the work being done by the medium during relaxation is used, the work required for the compression stage and a mechanical output power are available as well as to continue the cycle by repeating all of these steps.

It is another object of the invention to provide an apparatus that utilizes thermal energy into mechanical energy a gaseous medium.

In accordance with this aim, a device according to the invention comprises the following parts: Compression devices for compressing the working medium in a ratio "r", heat exchanger outlet devices for supplying heat to the compressed medium at a constant pressure, means for supplying further heat with constant volume, expansion devices for expansion of the compressed and heated gas in an expansion ratio "E", where "R" is greater than "r", means for connecting the relaxation and compression means to derive a first part of the relaxation devices for the converted energy for a drive of the compression devices, heat exchanger input devices, those with the heat exchanger external devices for the purpose of

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a withdrawal of. Heat from the expanded gas and providing this heat for the heating mentioned above are operationally connected at constant pressure, cooling devices for the extraction of further heat, devices to divert a second part of the energy converted in the expansion devices into mechanical energy Exit, and facilities for each consecutive periodic operation of the above devices.

fc The invention is illustrated below with reference to the drawing, for example described; in this show:

Fig. 1 is a circuit diagram of a simple, known gas turbine power plant and the computer calculated thermal efficiency as a function of the pressure ratio,

Fig. 2 is a similar circuit diagram as in Fig. 1, which as Difference includes a reheating and regeneration facility in the power plant,

Fig. 3 the basic components for the known Rankine cycle w process,

Figures 4A, 4-B, 4C, diagrams of the temperature-entropy

Characteristics of the Rankine cycle when applied a wet, dry or ideal relaxation,

FIG. 5 is a circuit diagram of a heat engine according to FIG Invention on its upstroke,

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6 shows a circuit diagram similar to that in Pig. 5, in which the Machine is shown on its downstroke

Fig. 7 is a pressure-volume diagram of the ideal thermodynamic Heat cycle according to the invention and

Figures 8 and 9 are graphs of computer ’computed performance characteristics of the ideal thermodynamic cycle according to FIG. 7 & it carbon dioxide "or air as Work media.

A conversion of thermal into electrical energy can, be accomplished in two ways;

1. Direct conversion

2. indirect conversion.

For use with Eadioisotope sources only need two methods of direct conversion to be considered, namely those with (a) thermoelectric and (b) thermionic Conversion systems. Both systems are operated and are in diverse development by a number civil and military organizations primarily in the United States. The newest and most comprehensive source of evidence the "Proceedings" of "Intersociety Conference on Energy Conversion Engineering" by 1967 · A series of power generation systems that are under the SNAP program - Systems for Nuclear Auxiliary Power - apply one or the other of two energy conversion processes at. The current "state of the art" of these systems is summarized in the table below shown.

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Characteristics

Energy conversion technology Thermoelectric Thermionic

Typical source and sink temp. S.

54-0 ° - 820 ° 0 (1000 ° -1500 ° J] to 65 ⁰ C (150 ° K

1650 ° C

to 650 ° C (1200 ⁰ F)

Overall efficiency

8 % - 15 %

Operational experience

A few years with a few energy generating systems low power

Limited laboratory experience

advantages

no moving parts; long, quiet operation

no moving parts; easy; low noise

disadvantage

low effective high temperature; Degree; heavy weights; difficult burning; Material production problems

It is unlikely that the conversion efficiency of these plants will be sufficient in the near future Degree is increased compared to the requirements discussed above. In addition, a Porsche and development work in these areas can be costly and lengthy. For these reasons, the direct conversion systems excluded for the present application.

The indirect conversion systems, which can also be called dynamic conversion systems, make use

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from a heat engine to thermal to mechanical To convert power that is subsequently easily converted into electrical power. There is a choice "re the thermodynamic cycle with which the heat engine works. The cycle processes of the previously built and the dynamic conversion systems under intense development are the following:

(a) Brayton cycle

(b) Rankine cycle

(c) Stirling cycle

Only "closed cycle processes" are used, since a cycle can only be "open" if air is used as the "working medium".

Conversion systems working with these circular processes become investigated below with particular reference to the subject of the application ·.

The Brayton cycle

A gas turbine power plant usually operates after this cycle. The working medium is a gas that does not change phase during the cycle. Any gas can be used as a Working medium can be used because the performance of the system is, theoretically, independent of the properties of the working medium. There are no degradation problems here of the working medium due to high temperatures or radiation.

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The thermal efficiency of the Brayton cycle is extremely dependent on the maximum cycle temperature sensitive to component performance. The basic circular process needs to be modified to make it reasonable to achieve thermal efficiencies. The changes include regeneration and rewarming.

A detailed analysis of the Brayton cycle with various modifications is done with the help of a digital computer been executed. The results of the analysis are shown in FIGS.

In Fig. 1 is the performance of the so-called OBTX circuit press shown. A schematic plan of one based on the CBTX cycle working system is shown therein. The working medium is compressed in the compressor 0 and after passage heated by the regenerator X in the heater B using a radioisotope source. The gas expands then through the turbine T, which exhausts through the regenerator Σ. Provided that air is used as the working medium an open cycle process is shown. The one on the turbine shaft The power available beyond the power requirement of the compressor is the useful power.

The following assumptions were made in deriving the in lig. 1 made results shown;

maximum cycle temperature »the maximum! temperature at the isotope source - 1570 ° R (600 ° 0); Sink temperature - 520 ° Rj

isentropic efficiency of the compressor - 80 % \ isentropic efficiency of the turbine - 85 % \ efficiency (efficiency) of the regenerator - 80%;

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Pressure loss through the heater - 2 % of the e-compressor delivery pressure; and

Pressure drop through the regenerator - 0.07 kg / cm (1 psi).

The partial efficiencies are assumed to be the uppermost limit of what can be expected within the current state of the art for the small sizes that come into being for the applications under consideration .86% at a pressure ratio of 2.6, which, when multiplied by the generator efficiency (about 85%) and the mechanical efficiency (about 90% ), gives an overall conversion efficiency of about 13 %.

The efficiency can be increased somewhat by using a more complicated system, as shown in! Fig. 2, which is referred to as the CBTRTX cycle. This cycle includes both rewarming and regeneration. The working medium is at the maximum cycle temperature after a partial expansion by the turbine T _x . reheated and relaxed again to the lowest possible pressure by the turbine Tp. Obviously, this arrangement is more complicated than the previous one.

The same assumptions as for the CBTX cycle were made when deriving the results for the in lig. 2 is made of the CBTRTX cycle shown by curve A. In addition, it has been assumed that the enthalpy decreases through the two turbines T ^ and Tg are the same as this has been found to give the best results. The maximum efficiency of 26 % can be seen at a pressure ratio of 3 »4 ·.

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That results after multiplication with the. Generator efficiency (about 85%) and the mechanical efficiency (about 90 ° / o) an overall conversion efficiency of nearly 20%.

The reason for the low efficiency is primarily the low value of the maximum cycle temperature 1570 ° H (600 ° C), which is assumed for the application in question is. This should be with the maximum cycle temperature of approximately 2400 ° R can be compared with that of modern gas turbine power plants work.

Another notable point is the extreme sensitivity of the Brayton cycle to partial efficiencies. That is shown by curve B of Fig. 2, in which the performance of the CBTßTX cycle process with slightly lower partial efficiencies is shown. The results of curve B are derived from the following assumptions:

isentropic efficiency of the compressor = 70 % \ isentropic efficiency of the turbines = 75 % \ efficiency (efficiency) of the regenerator = 75% i all other assumptions correspond to those given.

The maximum efficiency of 10.96 % occurs in this case at a pressure ratio of 2.6. This extreme sensitivity of the Brayton cycle to partial performance requires very careful adaptation of the various system components for optimal performance. Long and careful development work is the result.

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For this reason, the Brayton cycle can be used for the application under consideration, in particular with regard to the advantages of the cycle processes discussed below can be excluded.

The Rankine cycle

The Rankine cycle is inherently very efficient because it is mainly a heat supply and heat dissipation takes place at the highest or lowest cycle temperature. The basic components of a Rankine system are the turbine, the feed water pump, the heat source and the heat sink or condenser shown in FIG. 3 are. The main feature of the system is that the working medium changes phases during the cycle is subjected.

Steam is almost exclusively used as a working medium in Rankine cycle systems been used. Unfortunately, it is entirely unsuitable as a working fluid in small units the size considered here. The eligibility or unsuitability of a fluid as the working medium for a Rankine system thereafter decided how it would be the design and performance of the turbine, which is the most critical component of the the entire system.

The influence of fluid properties on a turbine design can be shown using standardized construction parameters as applied by Balje. These are

1/2

the (specific) nominal speed N =

and the (specific) nominal diameter D _g = -

1/4

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where Ή = the revolutions per minute,

Q = the volumetric outlet flow,

D = the turbine diameter and

H. = the isentropic heat removal by the turbine

indicates.

BaIje has contours of constant efficiency over N ^ - D ₁ ^

s s

Points indicated. A turbine then has to work with fixed values of Ή and D for optimum performance. Hence, FxD = constant for this point, so that holds

= constant

H.
xs

This shows that the peak speed of the turbine is directly proportional to the square root of the decrease in enthalpy due to the turbine is.

Both from a mechanical and aerodynamic point of view hence it is desirable to keep the peak speed of the turbine low. This then requires a slight decrease in enthalpy of the working medium between source and sink temperature. At this point the fluid properties their influence.

The three important properties of the fluid that have a significant impact on turbine design and performance have are:

1. the critical pressure,

2. the molecular weight, and

3. The nature (property) of the vapor after undergoing an isentropic expansion from a saturated one

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State is subjected to; that is, whether it becomes "wet" or "overheated" after expansion.

The influence of the critical pressure on the fluid can be seen from the following approximation; The relaxation through the turbine is, ideally, isentropic. The source and sink temperature and the mean Pressure levels in the system are assumed to be fixed.

We have:

411 ρ ■ ° *

In O

ι InTr vlnVr

Z «Compressibility I factor =

PV

Λ InZ = InP + In? - InR - InT

from which we get:

InTr]

a InVr \ -4 d InTr

Pr

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In the above relationships are:

Vr = relative volume = V / Vc, where Vc is the critical one

Volume is
Tr β relative temperature = T / Tc, where Tc is the critical one

Temperature is, Pr = relative pressure = P / Pc, where Pc is the critical pressure

is and
Cp = spec. Warmth at constant pressure.

From the tables, which give the properties of the corresponding states (change of Z with P and T), one can see see that the L H.S. of equation (2) with an enlargement of P increases. In other words,

BInVrJ Pr

decreases as P increases. From this it can be seen that a low critical pressure of the working medium is desirable.

The influence of the molecular weight of the fluid is from the the following evidence:

The energy of a unit mass of steam that is expanded between two fixed temperatures is approximately reversed proportional to its molecular weight, since the enthalpy loss = Cp (T ^ -T ^)

is,

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whereby:

R = the universal gas constant; K = the ratio of the spec. To warm; J = the mechanical heat equivalent; M = the molecular weight;
T ₁ = the source temperature and
Tp = the sink temperature

indicates.

It follows that for a given turbine output, the mass flow density is directly proportional is the molecular weight of the working medium. The volume flow density, however, remains unchanged. The speed of the fluid after expansion through the turbine nozzles is inversely proportional to its molecular weight. This works out out of the relationship

Speed =

-T * "TT ( ¹ ~

1/2

1/2

The speed of the turbine therefore increases with the increase in the Molecular weights of the fluid. Since a defined volume flow density is maintained for a given output power should be, the components of the turbine can be larger and more effective in power plant units smaller Performance to be made.

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The other essential property of the fluid in turbine design is the slope of the saturated steam line in the Temperature entropy coordinates. If this inclination, as with Water and liquid metals, is negative, the steam becomes "wet" after the expansion through the turbine nozzle, from which Erosion problems arise. When the slope is positive, as is the case with many organic fluids, the vapor after the Relaxation overheats and has to be cooled down from overheating before it can be condensed. Ideal is of course fe an infinite slope of the line. These three cases are shown in Figure 4-.

These considerations show that the use of a fluid with a low critical pressure and a high molecular weight, especially with the considered very small turbine sizes, is desirable. So there is steam from these unsuitable for both reasons.

Much work is being done on the use of mercury as a. Working fluid has been used for small Hankine cycle power plant systems. The advantage of its high molecular weight (200) seems to be more than offset by its high critical W see pressure (200 atm). Mercury also has a "wet" relaxation. It also introduces many additional difficulties in terms of the mechanical construction of the turbine components. These arise mainly from the difficulty of lubricating the bearings with mercury and from the extreme corrosive effects of mercury, especially when both liquid and vapor are present. The SNAP 1 power plant system, which used mercury as the working fluid, had the following performance characteristics:

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Electrical output power = 500 W "at 115 V; turbine inlet temperature = 704 ° C; condenser temperature = 167 ° Gj
Turbine speed = 40,000 rpm;
Isentropic turbine efficiency = 46% \ AC machine efficiency = 86%; Overall efficiency of the system - 9, M-% and service life = 104 days.

Another small Rankine cycle power plant system with mercury as the working medium is being intensively developed at Junkers Plugzeug- und Motorenwerke AG, Germany.

From all of the above, many organic fluids appear as working fluids for small ones Rankine cycle power plant systems particularly come into question. Monochlorobenzene, Dichlorobenzene, FC-75, etc. are typical examples of such fluids, usually these fluids have low critical pressure, high molecular weight, and a "dry" vapor deposition line. Their thermodynamic properties, their thermal and radiation stability have not yet been fully investigated. Nonetheless, there is little doubt that they are ideally suited for applications such as that contemplated here are.

Small eankine cycle power plant systems with an output in the range of 200 to 800 W, the organic working fluids apply have been put on the market. These are main used in conjunction with solar energy sources

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been. A typical design of such a 400 We unit is as follows:

Boiler temperature = 120 ° C;
Condenser temperature = 55 ° C;
thermal braking efficiency = 10 ° / o \ generator efficiency = 75 ° / ° and
Overall efficiency = 6%.

The adaptation of such a unit for the application under consideration can be the subject of a development study. An obvious problem is the thermal stability of organic working fluids at elevated temperatures.

The Stirling cycle

Although the conception of the Stirling cycle dates back to 1816 backdated, the development of a heat engine operating with this cycle has only recently been successful been carried out to the end.

The Stirling cycle is a regenerative cycle and achieves, in theory, the highest possible thermal efficiency between a given set of source and sink temperatures. The thermal efficiency of the Stirling cycle is the same as that of the Carnot cycle. The development of one with this thermodynamic cycle working mechanical device has required a very ingenious and sophisticated mechanical design. the Philips machine is a very complex mechanical arrangement with two coaxial pistons, which are in a cylinder without

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Lubrication go back and forth. The working medium is either Helium or hydrogen gas at an average pressure of about 110 atmospheres. These are high mean pressures required to obtain a reasonably high specific output power (power / weight) from the machine, and naturally pose problems of dynamic sealing between the piston rod and the cylinder. the maximum source temperature is currently limited to 700 ° 0. It is intended that this will be at 800 ° 0 and the mean gas pressure to 220 atmospheres when using materials with better fatigue strength in the vicinity To increase the future.

The performance data given for a commercially available machine are indeed very impressive. she be:

Source temperature »700 ° C;

Sink temperature = 25 G;

thermal braking efficiency «41% and total efficiency including the electr. Generator (85 %) »35% ·

The Carnot efficiency for these temperature levels is 69 * 4 %. This shows that this machine achieves a relative efficiency (actual efficiency / ideal efficiency) of approximately 60 %.

The adaptation of the Stirling engine to an isotope source could be a goal stroke for future developments, however the license costs, the operational safety and the materials constitute problems.

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The invention

It will be a simple mechanical one, with a new thermodynamic one High-efficiency cyclic process device proposed to be connected to a cadioisotope source can be customized. Although the efficiency of this thermodynamic cycle is not as high as that of the Stirling cycle, it is higher than that of most known cycles. The simplicity of the appropriate mechanical device may make the concept of this heat engine seem worth such a development as run a long-term project.

The machine consists of a reciprocating mechanism that feeds the working medium to the various in FIGS and 6 processes shown. The ideal thermodynamic cycle for the working medium is shown in FIG. 7 shown. The following points should be noted in connection with this machine.

(a) The expansion ratio is higher than the compression ratio. This results in using the top of the indicator chart, which is normally in known machines where the compression and relaxation ratio has been lost is approximately the same. One can achieve this because the compression and expansion processes are on opposite sides Side of the piston of the machine according to the invention take place.

(b) The regeneration process at the end of the compression process (at low compression ratios) uses some of the energy of the gases that have relaxed.

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This increases the mean temperature at which heat is supplied and the mean temperature lowered, with the heat is dissipated in the cycle.

(c) The process of supplying heat takes place essentially at constant volume in the time interval which is required for the entire upstroke of the piston is.

It is well known that all of these factors contribute to a high thermal Efficiency of the thermodynamic cycle. It is nevertheless correct that the entire working medium does not complete the ideal thermodynamic cycle shown in Fig. 7. Furthermore, there are imperfections due to component parts inevitable losses. Nonetheless, it is possible that the relative efficiency (actual efficiency / ideal efficiency) of this machine comparable to that of the Philips Sterling machine achieved in terms of their comparatively simple mechanical design.

A detailed analysis of this thermodynamic cycle has been carried out with the aid of a digital computer. The maximum cycle temperature was 600 ° C (1570 ° E) fixed as stated. The following gases have been restored to their capacity as working media in the Machine examines: air, nitrogen, carbon dioxide, hydrogen, helium, argon and sulfur dioxide. Air and carbon dioxide perform better than the rest, with carbon dioxide giving the best overall performance.

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The computer-calculated results with carbon dioxide and air as the working media are plotted point by point in FIGS. 8 and 9, respectively. The regeneration process can only be carried out as long as the temperature at the end of the compression process is lower than the temperature at the end of the expansion process. Curves A and B in Figures 8 and 9 therefore merge "at any value of the compression ratio. Curve 0 in Figures 8 and 9 indicates the specific output power of the machine; the output power per unit rate of the mass flow density W of the working medium through the machine. Thus, for a given output power, there are two ways to keep the dimensions of the machine at any desired value:

(a) change in engine speed;

(b) Change in the mean pressure and consequently the mean density of the working medium in the machine.

A practical limit to the minimum size of the machine may be given by the speed, which is limited by the heat transfer rate in the regenerator. The minimum mean pressure can also be determined by the heat transfer.

An analytical expression for the thermal efficiency of the ideal cycle is derived in Appendix A. Some of the performance calculations for a carbon dioxide machine for this application are given in Appendix B.

These preliminary performance calculations show that the new heat engine concept given here is indeed a of great importance as a prime mover for many applications

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has and for further research and development work "justified. Achieve a high level of thermal efficiency the following advantages:

(1) a high power / weight ratio;

(2) fewer noise and pollution problems;

(3) the use of any type of heat source; nuclear sources, fossil fuels or the like.

It also shows that this circular process can be carried out with simple mechanical components can be realized, which simplifies the development.

Attachment; A.

A unit mass of the working medium is assumed. In the ideal cycle, the regenerator has an efficiency of 100 %.

It is

and

now follows: (T ₅ ZT ₄ )

CVV.

-1

P ₅ ZP ₄

. P ₅ ZP ₄ «P ₁ ZP ₂ . T ₅ ZT ₄

(T ₅ ZT ₄ )

P ₁ ZP ₂ - (T ₁ ZT ₂ )

₁ ZP ₂

₁ ZT ₂

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where CR = the compression ratio and

V = the ratio of the specific heats C / C

indicates.

The heat supplied during the cycle is C (0

P The heat dissipated during the cycle »

Cp (T ₂ -T ₁ ) Op-T ₁ T (CE) ^ - ¹ -1J

thermal efficiency

_Λ dissipated heat _{m 1} Cp. ^J 1 ß ^0R) " ¹ J ~ supplied heat '" ov.T H- (CR) ^ ^ ¹ "^

This relationship indicates the thermal efficiency of the cycle in the form of the compression ratio, the maximum cycle temperature T ₄ and the minimum cycle temperature T ₁ .

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Attachment; B. Carbon dioxide machine;

Pressure ratio: 3 »

Expansion ratio: 4.1 j thermal efficiency: 50.9 % (FIG. 8); specific output power: 44.159 W / lbm per second (Fig. 8); relative efficiency: 60% (provided); Generator efficiency: 85% (provided) and total conversion efficiency: 50.9 x 0.6 χ 0.85 ^β 26%.

Actually desired output power »600 W

600 specified ideal output power » ^β ^ 75 ^, molecular weight of carbon dioxide» 44.01, density of carbon dioxide at 540 ° R and 1.015 kg / cm ² (14.5 psia)

; Volume flow density through the machine per second

assumed machine speed «1000 rpm

Displacement in the suction chamber per stroke ■ 0.242 χ60 x 1728 _{β 25 1 inch} 3

TOü ^{d :?} ' ^inc

assumed stroke / bore ratio for the machine is 1.0 β 25.1 where D ■ is the cylinder diameter

/.D a 8.05 cm (3.17 ") length of the working cylinder * 17.8 cm (7 ^W ) If desired, the machine can be made more compact by increasing the machine speed and / or the mean pressure of the carbon dioxide in the machine .

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Claims (6)

Claims / 1.! Process for converting energy available in a thermal source ^ - into mechanical energy in a continuous, closed cycle sequence using a working medium which is gaseous at all times, characterized in that the working medium is in a first space is compressed with a lower mean temperature and a lower mean pressure in a compression ratio "r", that a heat H is supplied to the compressed medium at a substantially constant pressure, that a further heat H is supplied from the source to the compressed medium at a substantially constant volume is supplied that the medium is expanded in a second space with a higher mean temperature and a higher mean pressure in an expansion ratio ¹¹ R ", where" S "is numerically greater than ^M r ^M , that an amount of heat nominally equivalent to H the relaxed medium is withdrawn at substantially constant pressure, this heat memenge H is available for supplying heat to the compressed medium at constant pressure, so that the medium at essentially constant /
υσϋ is cooled that the work done by the medium during the expansion in the second space is used firstly as work required for the compression in the first space and secondly as a mechanical output power and that all the preceding steps are repeated periodically. 009835/1267 2. Proceed according to. Claim. 1, marked thereby e t that the compression and relaxation steps are nominally adiabatic. 3. The method according to claim 1, characterized marked eichn e t that the first and the second space together form a fixed volume, the respective spaces through a movable member are separated within the fixed volume and that the net force acting on the movable member is effected at any time by the pressures acting on its opposite sides. 4-. Device for "converting energy available in a thermal source into mechanical energy using a gaseous medium, characterized in that it has compression devices for nominally adiabatic compression of the working medium in a compression ratio" r ", heat exchanger output devices for supplying heat to the compressed medium at essentially constant pressures, devices for supplying further heat to the compressed medium at essentially constant volumes, relaxation devices for nominally adiabatic expansion of the compressed and heated medium in an expansion ratio "R", where ¹¹ R "is numerically greater as "r", devices for connecting the Ent tension and compression devices for dissipating a first part of the energy converted in the expansion devices for a drive of the compression devices, heat exchanger input devices, the m it the heat exchanger outlet devices for the purpose of extracting heat from im 009835/1267 substantially constant pressure from the relaxed medium and one of the available at substantially constant pressures are operationally connected to the compressed medium supplied heat, cooling devices for a withdrawal of further heat at substantially constant pressure from the relaxed medium, devices for Deriving a second part of that in the expansion devices converted energy to a mechanical output, and facilities for successive, ) has periodically repeated actuation of all preceding devices. 5. Apparatus according to claim 4, characterized in that that the compression and relaxation means comprise piston and cylinder means. 6. Apparatus according to claim 5 »characterized in that that a single cylinder and piston are common for the compression and expansion devices wherein the cylinder includes intake and exhaust valves. "7 · Device according to claim 6, characterized in that that the piston comprises a piston rod forming the mechanical output. 009835/1267 Blank page