rexresearch.com
Edward PRITCHARD
Steam-Power Automobile
Steam-Power Automobile
http://blog.hemmings.com/index.php/tag/edward-pritchard/
1963 Ford Falcon: Steam Power Edition
by Mike Bumbeck
by Mike Bumbeck
In a perfect automotive world, history reveals that most of the technology we think of as new is quite old. Electric cars? Ask Detroit Electric (1907-1939) how that worked out. Hybrid cars? Dr. Porsche (1875-1951) himself could have expounded at length. Steam powered the Aeolipile, or Hero engine, 2,000 or so years ago in Greece, the industrial revolution few years later, a few automobiles along the way, and speeding locomotives up until not even that long ago. Only Superman could fly faster. Modern locomotives are series hybrids: a diesel engine powers a dynamo, which generates electricity to turn the wheels.
As combustion is better controlled at constant engine speed, the locomotive churns economically across great distances. Taking a series hybrid and a steam engine into a car was Australian engineer Edward Pritchard, who built a steam-powered 1963 Ford Falcon in the early ’70s. The Green Stripe Pritchard Steam Car efficiently burned most any combustible liquid, and drove about just like any other 1963 Falcon. While the Hemmings wood pellet fired 1977 Mercury Bobcat steam wagon is still in the developmental stage, this video reveals a steam powered Falcon that actually worked :
Ted Pritchard designed and built steam engine powering a 1963 Ford Falcon. Some vintage footage of Melbourne traffic as well.
https://revslib.stanford.edu/item/by200kt8858
http://www.virtualsteamcarmuseum.org/makers/pritchard_steam_power_pty_ltd.html
Pritchard Steam Power Pty. Ltd
http://trove.nla.gov.au/work/167572049?redirectedFrom=12815361&q&versionId=182636766
Herald & Weekly Times ( ca. 1968 )
Edward Pritchard inspects the steam engine he has designed with his father. Car is a 1963 Falcon.
https://s-media-cache-ak0.pinimg.com/736x/20/3e/4e/203e4eb59bd64585af304de30e696cd3.jpg
The Age ( 9 March 1972 )
US2008184944
Water tube boiler
Water tube boiler
A steam generator for generation of steam in a water tube boiler having first and second upright headers (16,18) in sealed communication with lower (12,14) and upper (13,15) inclining banks of tubes communicating therebetween. An end portion of the tubes in the upper bank (13,15), changes to a declining angle toward its communication with the upright header (16). The declining angle provides for increased separation of steam from hot water in the tubes.
[0001] This application claims the benefit of U.S. Provisional Patent Application No. 60/720,210, filed Sep. 23, 2005.
BACKGROUND OF THE INVENTION
[0002] 1. Field of the Invention
[0003] The invention disclosed and described herein relates to steam generators. More particularly the apparatus and method of employment herein disclosed relates to an improved design for a water tube boiler and steam generator which provides for improved separation of steam from residual water and enhanced protection from overheating of water tubes. The unique inclined design with curved end portions can be employed in any number of fields using steam including driving steam engines, for process steam, for steam heating, for hospital sterilizers, for most commercial power plants, for nuclear generators using steam boilers, or in any application where steam is employed.
[0004] 2. Prior Art
[0005] Water-tube style boilers for steam generation have been in use for decades and generally consist of natural-circulation style and submerged style water tube boilers. Water tube boilers were developed to satisfy the demand for large quantities of steam at pressures and temperatures far exceeding those possible with fire-tube boilers.
[0006] Water tube boilers have a low risk of disastrous explosion compared to fire box boilers or fire tube boilers, and they are space saving. They also provide for rapid steam raising and ease of transportation. However, water tube boilers have required that supply water should be substantially pure and specially treated to protect the steam tubes and may require special maintenance procedures for this reason.
[0007] Because of their safety and large production capacity for steam, water tube boilers are employed in products from steam engines to nuclear power plants and are considered an especially safe design for steam generation in a steam powered system. A wide variety of sizes and designs of water tube boilers are used in power stations, nuclear reactors, ships and factories. Well known designs such as those by Babcock and Wilcox have been in use for decades and those skilled in the art will understand the positioning and employment of the included water tube device herein, in proper communication with a heat source, for use in all such boilers.
[0008] Heating the water tubes of a water tube boiler or steam generator requires that fuel is burned inside a furnace, creating hot gas. The hot gases are communicated to the water tubes in various ways known in the art to heat up water in the steam-generating tubes.
[0009] Submerged water-tube boilers generally employ a means to heat water or fluid in the steam generator. The heat from fossil fuels, nuclear power, natural gas, or other sources, is communicated to a lower bank of inclined tubes through a first substantially upright header. The first or lower bank of tubes is inclined to communicate steam upwards through a plurality of the vertical headers. In such submerged boilers, the lower bank of tubes is substantially submerged in the heated water being communicated from the first upright header. Each of the lower bank of tubes communicates at an inclined end with a second substantially vertical header wherein steam rises in the second header and water will return to the reservoir below feeding the first header.
[0010] An upper bank of tubes communicating with the second header above the water line, receives the steam communicated through the second header from the lower bank of tubes, and communicates that steam through the upper bank of tubes at an inclined angle from the second substantially vertical header back to the first header. A preferred inclining angle for the first and second bank of tubes is at an angle between 11 and 15 degrees with a current especially preferred mode being substantially 12 degrees.
[0011] Various patents such as U.S. 309, 282, (Babbitt) describe such conventional submerged water-tube steam generators and all suffer from inadequate separation of remaining water from the steam which has been communicated to the upper bank of tubes. As such, there exists a need for an improved water-tube style boiler or steam generator which both dries and separates water from the steam. Such a device should also minimize the danger of overheating the water tubes which damages the apparatus and in doing so, results in an increased power rating for the steam generator device. Such a device should provide steam for turbines and the like which is substantially free of water droplets which can severely damage turbine blades.
SUMMARY OF THE INVENTION
[0012] The disclosed device and method of forming the device provide for an improved water-tube boiler or steam generator, which overcomes the above-noted deficiencies of prior art. The disclosed device is suited for use wherever water tube type steam generator devices are employed in combination with a properly communicated heat source to produce steam whether it be a liquid or gas communicating the heat from a heat source to the water tube boiler.
[0013] The device features water tubing which is divided into two sections or banks. A lower section features a plurality of tubes each of which angle upward from a first end, which is in sealed engagement with a first vertical header. Each of the plurality of tubes in the lower section is in sealed engagement at the upper end, with a second substantially vertical header. In one mode of employment, the device is in operative communication with a heat source in the form of hot gases from a furnace. In other modes of employment, the device may be employed with the entire lower tube section, submerged in water as a submerged water tube boiler.
[0014] In operation, heated water is communicated into an upright first header and thereafter into the inclined tubes of the lower section of tubes. Steam, and the hottest portions of water from the lower section of tubes reaching the axial passage of the second upright header, will naturally rise in the second header where it is thereafter communicated to a second bank of inclined tubes in sealed engagement between the axial cavities of the second header and first header.
[0015] The second bank of tubes is also angled upward from a lower end engagement with the second header to an upper sealed engagement of the opposite end of each tube, with the first header. Steam and/or water communicated from the lower tube section into the second header is thereon communicated into the tubes making up the second bank of inclined tubes where it will naturally rise toward the upper end of the first header.
[0016] Thus, the device features two banks of tubes, with all of the tubes of the lower bank or section angled upward from a respective starting end to respective termination ends at the second header. All of the plurality of tubes in the upper bank angle upward from starting end in sealed communication with the second header, to their termination-in sealed engagement with the first header. The upper or second bank traverses the distance between the first and second headers in the opposite direction as those of the lower bank.
[0017] In the preferred embodiment of the device, at the upper end portion of each tube member of the upper bank of tubes, adjacent to their individual engagement points with the first header, every tube is curved to angle downward to its sealed engagement with the second header. Consequently, an upper end portion of each tube in the upper bank of tubes changes direction from an upward angle to a downward angle just adjacent to a sealed engagement point with the first header.
[0018] Currently, this change in the angle of the upper ends of the tubes making up the upper bank changes around the curve from the noted upward angle to a declining angle. A current preferred angle of the upward incline is substantially 12 degrees relative to the substantially perpendicular second header to a declining angle of between 20 and 30 degrees with approximately 25 degrees being the especially preferred angle at their juncture with the substantially perpendicular first header.
[0019] The change in direction resulting in a downward or declining approach of the upper end portions of the tubes making up the upper bank of tubes has been found to provide an excellent increase in the efficiency of the device in separating water from steam which is to be communicated from the upper end of the first header to the device requiring the steam. Steam in the pipes of the inclining tubes of the upper bank of tubes naturally rises toward the top of each inclining tube. Consequently, at the point at the upper end of each tube where the direction or angle of the tubes changes from an incline to a decline toward the second header, steam is separated and accelerated into the first header in an upward direction. The water portion of the mixture which is already on the lower half of each tube, continues down the declining slope of the tubes entering the first header. This bifurcation of steam and water achieves an extremely high degree of separation of steam from water not heretofore provided by the simple horizontal or inclining tubes of prior art.
[0020] It is therefore an object of the present invention to provide a water tube component for a water tube boiler which provides increased boiler efficiency and steam generation which can be employed in all types of water tube boilers using a heat source generating steam for power.
[0021] It is a further object of this invention to employ downward curved portions of substantially all upper tubes of the water tube component to achieve increased separation of steam communicated to a device requiring it, from water.
[0022] These together with other objects and advantages which become subsequently apparent reside in the details of the construction and operation of the invention as more fully hereinafter described and claimed, reference being had to the accompanying drawings forming a part thereof, wherein like numerals refer to like parts throughout.
[0023] With respect to the above description, before explaining at least one preferred embodiment of the invention in detail, it is to be understood that the invention is not limited in its application to the details of construction and to the arrangement of the components or steps set forth in the following description or illustrated in the drawings. The various apparatus and methods of the invention are capable of other embodiments and of being practiced and carried out in various ways which will be obvious to those skilled in the art once they review this disclosure. Also, it is to be understood that the phraseology and terminology employed herein are for the purpose of description and should not be regarded as limiting.
[0024] Therefore, those skilled in the art will appreciate that the conception upon which this disclosure is based may readily be utilized as a basis for designing of other devices, methods and systems for carrying out the several purposes of the present buoyancy engine. It is important, therefore, that the objects and claims be regarded as including such equivalent construction and methodology insofar as they do not depart from the spirit and scope of the present invention.
[0025] Further objectives of this invention will be brought out in the following part of the specification wherein detailed description is for the purpose of fully disclosing the invention without placing limitations thereon.
BRIEF DESCRIPTION OF DRAWING FIGURES
[0026] FIG. 1 depicts a view of the water tube apparatus herein described showing the improved configuration for use in as a segment of a water tube boiler or steam generator and adapted for engagement with a heat source to generate steam.
[0027] FIG. 2 depicts a view of the device of FIG. 1 employed as a submerged water tube boiler, showing angles of incline of both banks of tubes, and the especially preferred downward angles of the upper end portions of the second bank of tubes. Also shown is the submerged lower bank.
[0028] FIG. 3 depicts the improved separation of steam from water in the fluid flow when the upper end portion of the tubes of the upper bank communicates in a downward angle at their engagement with the first vertical header.
DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS OF THE DISCLOSED DEVICE
[0029] As depicted in FIGS. 1-3, the device 10 herein provides a steam generator or water-tube boiler which is adapted for operative engagement with a heat source such as a conventional furnace or other means for communication of heat to the device 10. Most such steam generators are formed of multiple segments of similar construction grouped to form a larger steam generator with the tubular components of the segments being substantially inline and parallel to each other. The device 10 with the aforementioned improved water and steam separation will provide significant improvement when used in any type of water tube boiler over the prior art. The device 10 adapted more mounting in operative communication with the chosen heat source to generate steam and features a plurality of tubes 12 and 13 for communicating steam and water through the device. The two inclining pluralities of tubes 12 and 13, are formed in two distinct banks.
[0030] A lower bank 14 features a plurality of tubes 12 which in the current mode are substantially parallel with each other, and having a fluid capacity sufficient for the intended purpose. Each of the tubes 12 of the lower bank 14 angle upward at an inclining angle "C" from a lower first end which is in sealed engagement with a first vertical header 16. Each of the plurality of tubes 12 in the lower bank 14 proceeds to a sealed engagement at an upper end, with a second, substantially vertical header 18. The first and second vertical headers 16 and 18 in the current preferred mode of the device 10 are substantially perpendicular to a level support surface, and parallel; however, it is anticipated that other angles for the vertical headers 16 and 18 to both the support surface, and each other, may be employed.
[0031] The device as shown in FIG. 2, in a particularly preferred mode may be installed as a steam generator in a submerged water tube type boiler configuration with the entire lower tube section submerged in water below the water level 19.
[0032] In operation for steam generation, heated water is communicated into the first header 16 and thereafter into the inclined tubes 12 of the lower bank 14 wherein steam and the hottest portion of water from the lower bank reaching the second upright 18 header will naturally rise in the second header 18. This steam and high temperature water is therein communicated to the second or upper bank 15 of inclined tubes 13 where it proceeds upward in the inclined tubes 13 from the second header 18 toward the first header 16.
[0033] The upper bank 15 of tubes 13 is angled upward at an angle of incline "D" from a first or lower end engagement with the second header 18 to a transition point (shown as line between "A" and "B") at a curve and then downward to a sealed engagement at a second end with the first header 16. Steam and/or water communicated from the lower tube bank 14 into the second header 18 is thereon communicated through the plurality of tubes 13 of the upper bank 15 where it will rise toward the second end engagement to the first header 16.
[0034] As noted, in an especially preferred mode of the device 10, which experimentation has shown to operate with improved efficiency, an end portion of each tube 13 of the upper bank 15, from a curve at a transition point adjacent to their respective individual engagement points with the first header 16, is angled downward in a declining angle "A" from a curved point along the transition point in each tube 13. This reversal in the angle at the upper ends of the tubes 13 of the upper bank 15 from the noted preferred incline to a declining angle or path in the end portion of each tube, has shown to provide unexpected results in steam and water separation and efficiency of the device 10. Currently the inclining angle of the tubes 13 yielding most favorable results when combined with the upright parallel first and second headers 16 and 18, is substantially 12 degrees relative to the substantially perpendicular second header 18. The declining angle of the end portion between the curved portion and the second end works very well at substantially 25 degrees heading toward the sealed engagement with the substantially perpendicular first header 16.
[0035] This improved efficiency in separating steam from water is yielded by a means for enhanced separation of water from steam being carried in the upper tubes 13 provided by the declining approach of the end portions of the tubes 13 at their sealed engagement to the upper portion of the first header 16. The improved separation of the steam and water in the tubes 13 provided by the declining end portion of the tube 13 is provided by the steam which rises toward the top of the tube 13 and the water on the bottom of the tubes 13 being accelerated during the decline. Steam in the tubes 13 at the sealed engagement to the header 16 already on the upper portion of the tube 13, is accelerated upward into the first header 16 as it reaches it. Water, which is already on the lower half of each tube 13 due to lower heat content and higher density, is also accelerated by the declining slope of the tubes 13 entering the first header 16. As the water is denser and being accelerated in a declining angle of velocity, it continues in the downward angle imparted by the end portions of the tubes 13 and into the first header 16.
[0036] The declining angle of the end portions of the upper bank of tubes 13 thereby results in a much hotter and drier steam being communicated into the upper portion of the first header 16 and onto the blades of a turbine, or for any other purpose requiring high pressure, dry, steam.
[0037] The method and components shown in the drawings and described in detail herein, disclose arrangements of elements of particular construction and configuration for illustrating preferred embodiments of structure of the present invention. It is to be understood, however, that elements of different construction and configuration, and using different steps and process procedures, and other arrangements thereof, other than those illustrated and described, may be employed for providing a steam generator or water tube boiler in accordance with the spirit of this invention.
[0038] As such, while the present invention has been described herein with reference to particular embodiments thereof, a latitude of modifications, various changes and substitutions are intended in the foregoing disclosure, and will be appreciated that in some instance some features of the invention could be employed without a corresponding use of other features, without departing from the scope of the invention as set forth in the following claims. All such changes, alternations and modifications as would occur to those skilled in the art are considered to be within the scope of this invention as broadly defined in the appended claims.
US3892502
Control of expansion ratio in rotary motors
Control of expansion ratio in rotary motors
A rotary motor driven by a pressurised working fluid such as steam or compressed air having a series of working chambers around the periphery of the motor, an inlet port and an exhaust port oppositely disposed at the periphery of the motor for the inlet and outlet of the working fluid and a further port or ports between those ports admitting further working fluid bled from the supply to the chamber past the inlet port but before the exhaust port to significantly increase the amount of working fluid in the motor at low speeds or high load.
BACKGROUND OF THE INVENTION
This invention relates to rotary motors of the type using steam or compressed air or other expansible gas or vapor. In particular the invention relates to the provision of a device for varying the expansion ratio of rotary motors of the type specified and which are not fitted with intake valves such as rotary inlet valves or inlet valves operated by means of link motions or cam shafts.
In a rotary motor or reciprocating engine which is used over a wide range of speeds as, for example, from zero to the maximum design speed, it is desirable to use a small expansion ratio or "late cut-off" of the high pressure working fluid in order to obtain more positive starting, high overload torque and better smoothness on starting. A smaller expansion ratio or "later cut-off" is also desirable, apart from when starting, in order to run against loads heavier than normal. Under normal loads, it is desirable to run on a higher expansion ratio (early cut-off) so as to obtain more economical use of the working fluid.
In some rotary and reciprocating engines, changes of cut-off are obtained, for example, simply by varying the arrangements of link motions or by changing the positions of cams or, on engines fitted with rotary valves, by changing the position of rotary valve sleeves.
However it is advantageous if intake valves can be eliminated and to rely solely upon porting of the rotary engine thereby reducing initial and ensuing maintenance costs.
SUMMARY OF THE INVENTION
This invention has for its principal objective to provide a rotary engine of the type specified in which the expansion ratio can be varied in accordance with load and speed requirements.
With the principal objective in view there is provided according to the present invention in a rotary motor driven by a working fluid introduced through an inlet port under pressure the improvements comprising, a bleed or bypass passage leading the working fluid to a later expansion stage in the motor, thereby increasing the amount of working fluid within the motor to do work.
Conveniently a pressure sensitive valve is provided in said bleed passage, the valve being sensitive to variations in pressure of working fluid in the bypass line and the main supply line leading to the inlet port. It will be understood that the external load on the motor is proportional to the degree of pressure of working fluid in the motor. With an increase in load the pressure of working fluid in the motor must increase to maintain speed.
The bleed passage may be restricted and provided in parallel to said valve to provide a continuous flow of working fluid which is particularly effective at low motor speeds to substantially increase the amount of working fluid and decrease the expansion ratio. Alternatively the bleed restriction may be incorporated into the valve construction so that even when closed the valve continues to pass working fluid to a later expansion stage.
Said pressure sensitive valve is conveniently sensitive to working fluid inlet or expansion stage pressure, or a pressure differential between inlet and expansion stage inlet pressures. When a differential is apparent the valve is opened to transmit increased quantities of working fluid to a later stage so that effects of later cut-off are achieved. The inlet pressure represents a datum pressure and the expansion stage inlet pressure is dependant for its value upon the load on the motor. Accordingly, at low speeds and high load a pressure differential will exist across the valve to open said valve. The datum pressure and thus the pressure differential between inlet and expansion stage inlet pressures may be further increased by manual control of the engine throttle, however, this aspect forms no part of the present invention.
It will be understood that according to the invention re-introduction of the working fluid may take place at several succeeding stages of expansion to a lower pressure in the motor.
A practical arrangement of the invention will be described with reference to an extensible vane or blade type rotary motor having variable volume chambers, with said blades being driven by steam or air, however, it will be understood that the invention can be applied to various types of rotary motors in which there is at least two stages of expansion of the working fluid. The arrangement is described having reference to the accompanying drawings which depicts a schematic form of the invention.
BRIEF DESCRIPTION OF THE DRAWING
FIG. 1 is an end sectional view on line I-I of FIG. 2.
FIG. 2 is a sectional elevation taken on line II-II of FIG. 1.
FIG. 3 is a sectional elevation showing a modified form of engine with a plurality of late admission ports for working fluid.
FIG. 4 is a partial sectional view of a modified pressure sensitive valve.
DETAILED DESCRIPTION OF THE INVENTION
There is provided a stationary cylindrical ported chamber 1. In sealing contact with said chamber 1 a plurality of blades 3 are provided which extend radially from bearing bosses 3b mounted on shaft 2 about centre 0 enabling said blades 3 to move about centre 0 independently of each other in the direction of rotation shown. The blades 3 are constructed with their inner portions and bosses 3b forked to fit inside each other along the shaft 2. The blades 3 pass through sealing segments 4a mounted in a rotatable cylindrical structure 4 which is integrally formed with an end flange 4c at the drive end and with a removable flange 4d at the opposite end. The blades 3 are stepped down at 3a in overall width (taken along the axis) to clear inner portions 4b of the end flanges of the structure 4. At the driving end of the structure 4 an output shaft 21 on centre C extends from the end flange 4c.
Shaft 2 is held into end wall 23 of chamber 1 by means of a nut 22. Output shaft 21 runs in bearings machined into an extension 24 of the opposite end wall of chamber 1.
The ports, inlet 6 and outlet 7 in the chamber 1 are placed approximately opposite one another.
It is preferred that a sufficient number of radial blades 3 are provided so that at least two expansion stages or chambers are formed before exhausting through the exhaust port 7.
Accordingly smooth running of the motor is obtained despite the lack of intake and exhaust valves as well as providing a greater expansion ratio.
Referring to FIG. 1 bypass or bleed fluid passes through a restricting orifice or passage 8 of a definite predetermined minimum control area in bleed line 9 leading from the main supply line 6a to a later expansion chamber supplied by inlet pipe 16. The amount of bleed fluid is increased by opening up an additional passage area controlled by means of a pressure sensitive valve 11 which may be diaphragm controlled and interconnects passage 10 to passages 9 and 15 through needle valve 14. The diaphragm 12 may be sensitive to the pressure existing at a datum point such as at the fluid intake (not shown) or to a pressure differential such as the difference between the pressures at the intake 6 and at the point of readmission to a later expansion chamber as at pipe 16 of the working fluid as shown in FIG. 1. In this arrangement the diaphragm 11 is subjected to the pressure in passage 10 (datum pressure) on one side and the pressure in passage 13 (chamber pressure) on the other side. The diaphragm is adapted to move in response to the bias created by the pressure differential whereby the needle valve is moved to open or close passage 15. If necessary a compression spring 17 may be provided to ensure that the needle valve 14 is closed at the appropriate time.
A major controlling factor over the maximum expansion ratio on the type of rotary motor illustrated is the number of blades 3. The larger the number of blades, the greater the expansion ratio obtainable.
It will be appreciated that the bleed fluid flowing in the restricted passage 8, 9 will have little effect at normal running speed, the volume of flow being of but a small proportion of the total flow of working fluid through the motor. Thus the bleed passage is mainly effective at starting and low speeds as desired. Also, while the restricting orifice or passage 8 is of definite predetermined controlling area, this is to be understood to be for a given set of circumstances relating to the input fluid pressure and relative size of the main supply line and the bleed or bypass line. Accordingly, it is understood that such restricting passage may be of a selectively variable type to achieve a predeterminable control area therethrough for different conditions.
Referring to FIG. 3, a modification of the embodiment described with reference to FIG. 1 is illustrated showing a plurality of bypass passages controlled by a pressure sensitive valve 11 feeding into more than one expansion stage. In this Figure similar reference numerals refer to like integers. The internal construction together with the function of the valve 11 is similar to that previously described with reference to FIG. 1 or alternatively similar to that described here-below with reference to FIG. 4.
The arrangement shown in FIG. 3 depicts explicitly multiple bypass of working fluid to two expansion stages enhancing the amount of late cut-off working fluid that may be supplied to the motor. Having reference to the first stage after inlet pipe 6 (reference A), two inlet pipes 16a and b are arranged to feed into this stage at the particular point in time represented by the diagram. The spacing of the inlet pipes 16a, 16b, 16c and 16d is selected by the designer to achieve optimum benefit from the late admission of working fluid. For instance, the spacing of inlet pipe 16c from pipe 16a provides that chamber A will receive working fluid from all three inlets 16a, 16b and 16c for an instant in time during its passage around housing 1. Similarly chamber B would receive steam from 16b, 16c and 16d for an instant in time slightly preceding chamber A. It will be understood the phasing or spacing of the bypass inlets may be chosen according to needs and the operating conditions of the motor provided always that working fluid is not admitted when a chamber is being exhausted through port 7. The valve 11a may operate in identical fashion to that already described with reference to FIG. 1. Namely, a diaphragm 12 is provided, which is subject to a pressure differential between working fluid pressure in expansion stages cut-off from the working fluid inlet pipe 6 and the working fluid pressure in the inlet pipe 6a. In the arrangement shown in FIG. 1 the pressure sensitive valve 11 is provided in parallel circuit with bleed restriction 8 in pipe 9 leading from inlet pipe 6a to a later expansion stage.
The pressure sensitive valve may be modified as shown in FIG. 4 by the provision of a non-closable valve and valve seat 14, 14a. All other parts of the valve 11a are identical in construction to that previously described with reference to FIG. 1. The seat 14a includes small slots or recesses 14b spaced therearound through which working fluid may pass even when valve element 14 is in engagement with the seat 14a. Accordingly, lifting of the valve element 14 off its seat merely allows for an increase in flow of fluid into line 15 and thence to a later expansion chamber. It is preferred but not essential that valve 11a be utilised in feeding working fluid to a plurality of later expansion stages as shown in FIG. 3.
US7536943
Valve and auxiliary exhaust system for high efficiency steam engines and compressed gas motors
Valve and auxiliary exhaust system for high efficiency steam engines and compressed gas motors
A steam engine with improved intake and exhaust flow provided by separate pairs of intake and exhaust ports located at both ends of a steam drive cylinder. A slide valve located adjacent to the drive cylinder provides for timed sealing of intake and exhaust ports during operation. Exhaust is facilitated by the provision of two paths of exhaust from the cylinder and the exhaust ports may be adjusted for a flow volume to meter exhaust steam flow to significantly reduce back pressure only at low speeds of said engine.
FIELD OF THE INVENTION
The invention relates to steam engines. More particularly, the invention herein disclosed relates to an improved design of the valve and auxiliary exhaust system for steam engines of both the double-acting and single-acting designs and in particular for uniflow steam engines with auxiliary exhaust. The design can be employed upon fixed timing engines or with added means for timing adjustment, upon variable timing steam engines.
BACKGROUND OF THE INVENTION
Single-acting and double-acting steam engines have provided power for industry and other uses for a long period of time. The single-acting steam engine may resemble a two and a four-stroke internal combustion engines in that a piston, connecting rod and crank are used per cylinder set. With the double-acting form of steam engine, straight line reciprocating motion is described not only by each piston, but also by each piston rod and crosshead. Motion is transferred from the crosshead via a connecting rod to the crank. The piston rod passes through a seal in the end of the cylinder and the steam is valved to work on the piston from above and also below it. This gives a “one stroke” action. With two double-acting cylinders, only four valves are required on a “full” uniflow engine of conventional design as against sixteen valves being required for an eight-cylinder four-stroke engine which exerts the same number of power impulses per revolution.
The uniflow engine exhaust system uses holes in the cylinder which are exposed to the top end of the cylinder adjacent to the piston near the bottom of its stroke. The same row of holes are exposed to the bottom or crank end of the cylinder adjacent to the piston near the top of its stroke. The length of the piston adjacent to the cylinder wall is equal or approximately equal to the stroke minus the diameter or length of the exhaust holes. (The exhaust holes in the cylinder can be seen in one of the photos on display). Clearance volume is provided at each end of the cylinder to allow for reasonable compression to take place at each end of a stroke.
A semi-uniflow engine is one in which exhaust valves are used to supplement the action of the exhaust holes in the cylinder wall. By employing the exhaust valves, the point at which compression begins on the return stroke of the piston can be delayed. Such an auxiliary exhaust feature is useful especially where exhaust is at atmospheric pressure rather than into a vacuum and/or, further, where compounding is utilized. Further, in single cylinder engines which are not necessarily self-starting, the auxiliary exhaust makes the engine easier to start. This is because it is easier before the admission steam starts the engine to rotate the engine against compression since with an auxiliary exhaust system compression acting against the piston begins later on the compression stroke.
In some early uniflow engines with auxiliary exhaust systems, the auxiliary or secondary exhaust steam traveled out through the same ports and passages through which previously admission steam entered. A disadvantage of this design is that the cooling effect of the exhausting steam lowered the efficiency of the engine. In other early semi-uniflow engines the auxiliary or secondary exhaust steam exhausted through special ports in the cylinder wall at positions between the main uniflow exhaust and the admission passages, the latter located near the cylinder ends.
Special valves such as poppet valves controlled these auxiliary exhaust passageways. These engines, if of the double-acting type, were fitted with four valves: two for inlet steam—one at each end of the cylinder, and two for auxiliary exhaust—one for the upper part of the cylinder and one for the lower part of the cylinder. A disadvantage of this design with its four valves plus the respective valve motions required for their operation is relative complexity. [See Skinner. P271. “Power from Steam,” R. L. Hills.]
PRIOR ART
U.S. Pat. No. 3,967,535 (Rozansky) while disclosing that the device relates to uniflow steam engines having a novel valving means for controlling the introduction of steam into the cylinders, is not concerned with auxiliary exhausting.
U.S. Pat. No. 3,651,641 (Ginter) discloses an engine system and thermogenerator therefor. Ginter in teaching a valving system seems primarily concerned with an internal combustion engine with water internal cooling and there are no uniflow exhaust ports and no auxiliary exhaust ports disclosed.
U.S. Pat. No. 3,967,525 (Rosansky), while disclosing that the device relates to uniflow steam engines having a novel valving means for controlling the introduction of steam into the cylinders, is not concerned with auxiliary exhausting.
U.S. Pat. No. 3,991,574 (Frazier) discloses a uniflow exhaust system in a rather complex structure. However, Frazier does not teach the employment of an auxiliary [uniflow] exhaust.
U.S. Pat. No. 3,788,193 (O'Conner) discloses a spool type slide valve for controlling both admission and auxiliary [uniflow] exhaust. However, O'Conner requires the employment of a complex system of powered cams to operate the disclosed valve. O'Conner teaches a complex double cam driven system, the cams having positive lift and drop as in “desmodromic” systems with complex chain drives to achieve variable valve timing. In the mid-position of the slide valve it appears that the admission and auxiliary exhaust ports are both closed. The variable engine “timing” or valve events are controlled by phase changes in their relative positions of the double cams and also with the angular displacements of the camshafts with the “variable” chain drive.
As such, there exists a need for an improved auxiliary exhaust valving system on steam engines with fixed timing which employs a simple mechanical operation to achieve the desired result. Such a device should utilize simple harmonic motion from a simple eccentric and should provide the required valve events by careful selection or design of the required bobbin admission and exhaust “laps” or the eccentric radius and also the phase relationship between the eccentric valve drive and the crank. Still further, such a device and system should be easily adaptable to a variable timing steam engines.
With this design, the valve events can be worked out using conventional valve diagrams, e.g. “Bilgrams Valve Diagram”. Still further, such a design should control the auxiliary exhaust in a manner similar to the conventional steam engine which exhausts through the common admission/exhaust ports. Employing such a control, the auxiliary exhaust should then be communicated through ports and passages separate from the admission passages which could be said to be in the correct uniflow tradition. Main central uniflow exhaust should also be utilized.
As can be seen and readily discerned by those skilled in the art, this invention can also be employed, if desired, to obtain variable valve timing using conventional valve gears such as Stephenson's link, Allan's link motion, Joy valve gear, Walschaert, etc. This enables forward and reverse operation plus changes of cut-off.
SUMMARY OF THE INVENTION
The disclosed device provides for an improved valve and auxiliary exhaust system when employed and yields a high efficiency steam engine or compressed gas motor. For fixed timing, as may be utilized for a stationary engine, a preferred embodiment utilizes movement for the slide valve in harmonic motion derived from a simple eccentric and connecting rod and obtains the required valve events by careful selection or design and inter-related functions of the required bobbin admission, and auxiliary exhaust “laps,” and the eccentric radius and the phase relationship between the eccentric and the crank. The slide valve would be adapted to move in a direction controlled by an eccentric set at between 90 to 180 degrees ahead of the crank controlling the piston. The design procedure for valve event timing is similar to that of a conventional non-uniflow outside-admission slide valve or other slide valve engine. Conventional valve diagrams such as Reuleaux's Slide Valve diagram can be used to assist in the design of this invention.
It should be noted that the device as herein disclosed shows employment for use in combination with a fixed timing engine for ease of illustration of the novel properties of the device and the great utility provided in steam or compressed air engines. However, those skilled in the art will no doubt realize that inclusion of a means to vary engine timing, such as a camshaft, could be added to the design disclosed herein, thereby providing a variable timing engine with improved efficiency, and all such modifications are anticipated to be within the scope of this invention.
The embodiments herein provide for forward and reverse control and change of cut-off to facilitate start-ups and obtain normal operation at more efficient early cut-offs as is usually required for a mobile engine. Again, conventional slide valve driving mechanisms such as Stephenson's Link Motion and Walschaert's Valve Gear can be utilized to drive the valves of this invention.
The embodiment further provides for built-in easy starting with low compression plus phasing out auxiliary exhaust under more load and speed. This is achieved with the area of the auxiliary exhaust ports designed so that, on start-up and at slow speeds, the steam flow is adequate to hold cylinder exhaust pressure close to exhaust pipe discharge pressure. This assists with easy starting of the engine and makes for smooth running at low speeds. However, with rising speeds and bigger throttle openings, more steam will pass through the engine. There will consequently be greater pressure drop through the auxiliary exhaust ports with the amount of pressure drop depending on the flow areas. The latter are designed to achieve the desired metering of the steam flows. This will lead to higher cylinder pressures and higher compression pressures, i.e., the engine will run more like a “full uniflow” type and higher efficiency can be realized.
The above design provides a device which is much simpler than alternative systems of controlling the extent and timing of opening of the auxiliary exhaust ports. This latter more complicated type of arrangement may be activated by devices sensitive to engine speed and/or amount of steam flowing through the engine.
The valve functions for controlling inlet steam and auxiliary exhaust steam are provided by a slide valve, preferably of the slide valve type. Slide valve designs were commonly used in conventional types of counter-flow steam engines, but in the device and method herein disclosed, the valve is used in a different manner in keeping with the uniflow principle. In keeping with that principle, one area of the valve controls one inlet steam function and a different area of the valve controls an auxiliary exhaust function. The two areas control steam flow through separate inlet and auxiliary exhaust ports and passages. Thus, the flows of hot inlet steam and the relatively cooler exhaust steam are kept apart. A single slide valve can be used to control both inlet and auxiliary exhaust steam in a single-acting engine and also in a double-acting engine.
In the device herein described and disclosed, the simplicity of the slide-valve design is retained. The slide valve may be driven by valve gear giving the valve simple harmonic motion or an approximation to it. The valve gear, as in conventional steam engines, may be designed to give reverse operation plus changes in cut-off. A valve and drive system similar to that described herein is suitable for use in an engine without central “uniflow” exhaust ports which are uncovered by the piston near the ends of its stroke. In this case, the “auxiliary” exhausts described herein will be the main exhausts.
With respect to the above description then, it is to be realized that the optimum dimensional relationships for the parts of the invention, to include variations in size, materials, shape, form, function and manner of operation, assembly and use, are deemed readily apparent and obvious to one skilled in the art, and all equivalent relationships to those illustrated in the drawings and described in the specification are intended to be encompassed by the present invention. Therefore, the foregoing summary is considered as illustrative only of the principles of the invention. Further, since numerous modifications and changes will readily occur to those skilled in the art, it is not desired to limit the invention to the exact construction and operation shown and described, and accordingly, all suitable modifications and equivalents may be resorted to falling within the scope of the invention.
Accordingly, it is the object of this invention claimed herein to provide a steam engine having an improved slide valve and drive cylinder design wherein one area of the slide valve controls one inlet steam function and a different area of the valve controls an auxiliary exhaust function.
It is another object of this invention to supply the disclosed steam engine wherein two areas of the slide valve providing improved operation control steam flow through separate adjacently located inlet and auxiliary exhaust ports and passages.
It is another object of this invention to supply an improved steam engine providing a slide valve control of overall operation which is much simpler than alternative systems of controlling the extent and timing of opening of the auxiliary exhaust ports.
These and further objectives of this invention will be brought out in the following part of the specification, wherein detailed description is for the purpose of fully disclosing the invention without placing limitations thereon.
BRIEF DESCRIPTION OF DRAWING FIGURES
FIG. 1 shows a sectional view of the device featuring a cylinder arrangement for a Uniflow, double-acting type steam engine.
FIG. 2a depicts the new design showing the engine of FIG. 1, depicting the valve laps of the present device.
FIG. 2b depicts prior art in the form of a valve design for a conventional non-uniflow steam engine and the conventional admission lap and an exhaust lap thereof.
DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS OF THE DISCLOSED DEVICE
FIG. 1, shows the cylinder arrangement for an embodiment of the device for employment with a uniflow double-acting steam engine including the herein disclosed and described novel valve design, which used in conjunction with a conventional harmonic valve drive mechanism, forms the main part of this invention. Not shown is the crankcase containing the connecting rod, crosshead, valve motion and other attendant parts of the engine for which the disclosed device is adapted for engagement. These latter parts may be conventional in design.
The main parts shown are the cylinder 1, the valve chest 2, and the piston, 3, which would best be fitted with piston rings (not shown). Also shown in FIG. 1 is the piston rod 4, and the slide valve 5, showing the continuous exterior surface to contact the valve chest and which also would be fitted with sealing rings but which are not shown. Number 6 depicts the cylinder head and the valve chest cap is shown as number 7. The piston rod sealing assembly is identified by number 8 and number 9 represents the valve rod sealing assembly.
In the preferred arrangement of the disclosed device as shown, the valve chest 2 is adapted for outside admission. Two steam inlet ports are shown as numbers 10 and 11 and the exhaust port is shown as number 12. The cylinder upper inlet steam passage 13 is shown at the upper area of the cylinder adjacent to the cylinder upper auxiliary exhaust steam passage 14.
In a central section of the cylinder 1 is the cylinder uniflow or central exhaust steam ports 15. At a lower end of the cylinder 1, is the cylinder lower auxiliary exhaust steam passage 16 and the cylinder lower inlet steam passage 17. It should be noted that steam inlet passages 13 and 17 and steam exhaust ports 14 and 16, while depicted as single passages, may be one or a plurality of passages to provide the volume communication required. It should also be noted that use of the terms upper and lower are for convenience sake and those skilled in the art will realize that positioning and operation of such engines is possible using different manners of positioning of the components described herein. Therefore the invention herein described and disclosed is employable for steam engines of any position and angle of operation.
In the position shown in the drawing FIG. 1, the piston 3 is shown as it would be moving downwards toward the lower end of the cylinder 1 and the slide valve 5 would be concurrently rising in the opposite direction of the piston 3. The upper part of the slide valve 5 at the face 18 is moving away from the piston 3, and has just cut off the inlet steam supply though the upper steam inlet 13 to the upper portion of cylinder 1 and to the top of the piston 3 which continues to travel downwards under pressure from the expanding steam in the upper portion of the cylinder 1. The length of the continuous side edge of the upper piston 27 of the slide valve 5 and the speed of the slide valve 5 in reciprocal motion to the piston 3 during each engine cycle to determine the length of time it will maintain this cut off of steam so long as it covers the inlet 13. Steam under the piston 3 in the lower portion of the cylinder 1 is exhausting through lower exhaust passage 16 and out the exhaust port 12, until cut off by the piston 3 moving downward wherein the piston's continuous side edge covers the exhaust passage 16. At this point in the timing of the device, compression of the residual steam in the lower portion of the cylinder 1, under the piston 3 will then begin.
When the piston 3 reaches the bottom of its stroke, steam above the piston 3 in the upper portion of the cylinder 1, will exhaust through the main uniflow or central exhaust ports 15. The side of the upper piston 27 of the slide valve 5 bounded by the lower face 19 will uncover its sealed engagement over the auxiliary upper exhaust port 14 which will vent exhaust steam also thereby emptying the upper portion of cylinder 1 through two routes of exhaust increasing efficiency of this operation. It has been found through experimentation that the total aggregate area of each set of the exhaust ports 14, and 16, may be adjusted to provide a means for metered steam flow such that the flow is adequate to significantly reduce back pressure only at low speeds, while at higher speeds and larger throttle openings, the engine will operate more similarly to that of a full Uniflow engine. This can be done through adjusting the sizes of the exhaust ports so that the volume of exhaust vented at lower speeds of the engine being built for use at desired speeds and loads has the desired reduced back pressure at the determined low speeds.
The slide valve 5 will uncover the lower steam intake 17 and allow admission of steam past rising lower face 20 at the bottom of the second or lower piston 25 opposite the upper piston 27, and communicate it to the underside of the piston 3. The piston 3 will then begin to rise from the force of the steam. In reciprocal action, the slide valve 5 with the first or upper piston 27 and the second or lower piston 25 operatively engaged by the valve rod 28 at an operative distance, now begins to descend, and the lower end slide valve 5 bounded by face 20 of the lower piston 25 passes the lower inlet steam passages 17 wherein the continuous side edge of the lower piston 25 seals the lower inlet steam passage 17 and causes cut-off of the steam communicated to the lower end of the cylinder 1. The duration of the cut off is determined by the length of the side surface of the lower piston 25 in the same fashion of the sealing operation of the upper piston 27 combined with the speed of the valve rod 28. The piston 3 is now moving upwards from the force of the steam in the lower end of the cylinder 1, and the slide valve 5 is concurrently descending in the opposite direction. This reciprocal cycle now continues similarly to that described above but for an “up” power stroke of the piston 3, rather than for a “down” power stroke.
Drawings FIG. 2a and FIG. 2b illustrate the improvement of the disclosed device and operation over conventional valve design for the disclosed uniflow engine by employment of auxiliary exhausts 14 and 16, compared with valve design for a conventional engine with outside admission slide valves as shown in FIG. 2b. Also shown in FIG. 2a is the unique admission lap 21 and uniflow auxiliary exhaust lap 22 of the disclosed device which is provided by the length of the continuous side wall of the piston at the upper cylinder 27 and lower cylinder 25. The continuous sidewalls of both the upper and lower cylinders of the slide valve 5, cover both their adjoining respective inlet and exhaust ports during each cycle for a lap period determined by the length of each of the two cylinders of slide valve 5 which are operatively engaged by the valve rod 28 therebetween and thereby define the admission lap 21 and exhaust lap 22 shown in FIG. 2a.
FIG. 2b to the contrary shows the close proximity of the admission lap 21 and the conventional exhaust lap 23 in a non-uniflow steam engine and the short duration therebetween and limitations on adjustment. Employing the discloed device, the full diameter part of each bobbin is extended so as to control exhaust steam flow separably through the auxiliary exhaust ports rather than through common admission and exhaust ports as in conventional designs. The disclosed design also does not require any special cams, chains, or valves in its connection of the slide valve 5 to the control system. The design may be carried out using conventional valve diagrams which incorporate specifications for eccentric radius and eccentric phase difference with the engine crank.
US3818699
FEED AND INJECTION WATER CONTROL FOR STEAM GENERATORS
FEED AND INJECTION WATER CONTROL FOR STEAM GENERATORS
A steam generator control system including a once through steam generator, a superheater thermostat sensing the temperature of superheater steam in the generator and controlling a fluid injection circuit connected in parallel to a portion of the steam generator coil to supply injection water to the coil and also, controlling the supply of feed water supplied to the steam generator, the feed water supply being controlled substantially proportional to the amount of exhaust steam issuing from a steam consuming apparatus together with extra water when called for by the thermostat, said control being provided by a positive displacement motor such as a rotary motor driven by the exhaust steam, and including auxilary governing means for the rotary motor to ensure more accurate proportionality between the speed of the rotary motor and the supply of feed water requirement to the steam generator so that full utilization of steam generator burner output is obtained for any given requirement.
This invention relates to the control of feed water and injection water flow into steam generators of the type known as flash boilers and once through steam generators having single or parallel coils. Fluids other than water can be used in similar "vapour" generators. It is pointed out, therefore, that "fluid" can be read for "water" and "vapour" for "steam."
One of the main problems in the development of compact steam generators as used for automotive steam engine systems is in the control of the generator. A large percentage of experimental steam car projects have failed because of the inability of the designers to solve the control problem. The aim is to obtain reasonably constant steam pressure and temperature at the outlet of the steam generator. During normal operation over a wide range of loads, control should not be at the expense of a reduction in burner output which causes undesirable reduction in steam generator pressure, in order to maintain safe temperatures throughout the steam generator.
The principle of the once through steam generator appears deceptively simple. Water is pumped in at one end and superheated steam is led away from the other end. A survey of the rather voluminous patent literature on the subject of control systems shows, however, that a wide variety of control "schemes" are proposed. It is clear to the applicant from extensive experimental trials and an appraisal of prior proposals that the correct control of a steam generator is not obvious even to those supposedly skilled in the art. Some of the problems involved are:
CONTROL OF THE BURNER
This is not a difficult problem. Quick response or feed back can be obtained with either pressure or temperature control.
CONTROL OF WATER SUPPLY
This is a more difficult problem. Although, for example, an increase in feed water supply will cause an almost immediate response on the feed water heating or economiser section of the steam generator in which the fluid is largely incompressible, there will be a delay in the feed back to a thermostat fitted to the steam generator in a steaming or superheater zone in which case there exists compressible fluid (steam) between the feed pumps and the thermostat.
SEVERAL PRIOR ART CONTROL SYSTEMS ARE REFERRED TO IN A GENERAL WAY:
a. Pressure Control of feed pumps (e.g., early White steam car). A disadvantage of this system was excessive blowing of the safety valve especially when delay in the pressure control caused too much water to be fed into the steam generator. The burner, under thermostat control, would endeavour to bring temperature back up, even at zero power output.
b. Temperature Control of feed pumps and water supply. Burner main control was usually a pressure type with an overriding high temperature cut-off. Some variations in temperature control systems as outlined in (b) are:
I. final Thermostat type. Control thermostat is situated in the superheater zone. Disadvantage -- too much response delay.
Ii. thermostat at end of evaporative zone. (e.g., British Patent 254,774, 1926, W. M. Cross.) Disadvantage -- Too much response delay.
Iii. thermostat(s) in evaporative zone. (e.g., U.S. Pat. Re 20045, 1936, J. Fletcher). Disadvantages -- Too much response delay. Also effected by inherent changes in boiling point temperatures with steam pressure changes. The latter applies in particular to automotive systems, where certain steam pressure changes take place in normal operation.
Iv. thermostat in feed water heating zone. Small response-delay with this system but thermostat is situated so far away from the final steam generator zone that poor control can result from secondary effects such as soot on generating coils modifying water-steam zone positions.
V. thermostat plus water injector. In British Patent Specification 568,722, 1945, M. H. Lewis states that up to 5 percent only of total feed water capacity is fed to the water injector nozzle in the superheater zone. Otherwise there is a danger of a high temperature peak before the injector point. See later for argument showing that this amount would be insufficient to result in good control but that, in some systems, additional increments of "base" water can be fed equal to the quantity of water injected.
There have been tried and proposed various combinations of thermostatic and water injection systems. Estimates are made later which show that a thermostatic and injection system alone cannot provide sufficient basis for a correct control but, from certain considerations, can be used to control up to only approximately 65 percent of the total water. An additional form of control must therefore be provided.
Early systems using variable capacity vaporising burners proportioned water and fuel. (e.g., -- Serpollet, later White steam cars.) Note that, with the type of vaporising burners used, roughly proportional air-to-fuel ratios were maintained. With modern pressure atomising burners, systems using variable fuel and air supplies are complex and relatively expensive, particularly as applied to small units.
There have been systems using auxiliary reciprocating engines or turbines driving feed water pumps (and other auxiliaries) in order to assist in matching water flow with burners demand. Some systems use water metering valves, sometimes dependent on hand adjustment. With some water injection systems, relatively large amounts of water, which are sometimes relatively cool, are injected into superheated steam causing thermal shock to the piping system. Thermal shock is a serious problem particularly under the difficult conditions encountered with the varying power requirements of an automotive steam power system where frequent operation of the injection control may be required. Thus with such systems it is undesirable to inject into the superheater zone.
Main engine-driven pump systems have the disadvantage that, at low speeds, particularly with a cold engine, the feed pumps pump insufficient water. Some prior systems are not fundamentally sound, in that they will not cope with a wide range of power demands. On some, to prevent local overheating, the burner is cut. This may reduce available power.
Applicants earlier Australian Patent No. 226,096, "Improvements in Steam Plants for the Control of Plant Auxiliaries proportional to the Steam Consumed," stated that, . . . "preferably the drive arrangement according to the invention is operated in conjunction with conventional temperature actuated means (thermostat) controlling a secondary feed pump, which is cut-in to boost the primary pump, responsive to changes in steam temperature within the steam producing unit." In practice, such conventional means did not prove adequate. Very considerable experimental and theoretical work was carried out before the control system according to the present invention was evolved.
As described in Applicants earlier U.S. Pat. No. 226,096, water quantities bear a direct relation to exhaust steam quantities rather than to burner rates. This means that steam generator water control can be largely independent of burner operation. Thus, boosting of the burner will not directly effect the water control system.
It is a principal objective of the present invention to overcome the abovementioned problems and provide a steam generator control system in which the quantitative components affecting the operation of the system namely the feed water pump means, feed water injection system and burner are controlled.
It is a further objective of the invention to provide a steam generator control system in which the auxiliaries are driven by an improved proportional exhaust steam motor drive in combination with a water injection system in which feed water flow rates and injector flow rates are controlled within certain proportions calculated empirically.
It is a further objective of the present invention to provide a steam generator control system in which known definite quantities of water are automatically fed into the base of the steam generator coils by proportionally driven feed pumps or controlled metering means and known definite quantities of water are injected (as required) into a known desired evaporator zone point of the steam generator coil, thus resulting in a fast response and stable control with minimum thermal shock at the injection point, and enabling full utilisation, under normal operation, of the burner output for a given steam generator capacity.
There is provided according to the present invention a steam generator control system comprising a steam generator, a burner, a once through coil heated by said burner supplying superheated steam to a steam consuming apparatus, a superheater thermostat disposed on the coil at or near the outlet end of the coil in proximity to said burner arranged to sense the temperature of superheated steam, a water injection circuit connected in parallel to at least a section of said coil and arranged to carry feed fluid by-passing said coil section to inject said feed fluid into a zone of the coil carrying fluid of higher temperature, said injection circuit including valve and metering means for controlling flow of feed fluid therein, feed water supply means arranged to normally provide feed water at a rate below the requirements of the steam generator and to intermittently provide an increased flow of feed water when there is a flow of fluid in the injection circuit said increased flow resulting in a total feed water flow in excess of the requirements of the steam generator, a positive displacement motor operated by the exhaust steam from the steam consuming apparatus and arranged to control at least said feed water supply means at a rate substantially proportional to the volume of steam consumed by the consuming apparatus.
The superheater thermostat may be positioned anywhere in the superheater zone of the generator. In another aspect of the invention there is provided according to the present invention a steam generator control system comprising a steam generator, a burner, a once through coil heated by said burner supplying superheated steam to a steam consuming apparatus, a superheater thermostat disposed on the coil at or near the outlet end of the coil in proximity to said burner arranged to sense the temperature of superheated steam, a water injection circuit connected in parallel to at least a section of said coil and arranged to carry fluid from by-passing said coil section to inject said fluid into a zone of the coil carrying fluid of higher temperature, said injection circuit including valve and metering means for controlling flow of fluid therein, feed water supply means arranged to normally provide feed water supply means arranged to normally provide feed water at a rate in the range of 60 percent to 90 percent of requirements of the steam generator and to intermittently provide an increased flow of feed water when there is a flow of fluid in the injection circuit, said increased flow resulting in a total feed water flow in the range of 120 percent to 180 percent of the requirements of the steam generator, the volume of fluid arranged to be injected by the injection circuit being in the range of 30 percent to 90 percent of the increase in feed water flow above that otherwise provided, a positive displacement motor operated by the exhaust steam from the steam consuming apparatus and arranged to control at least said feed water supply means at a rate substantially proportional to the volume of steam consumed by the consuming apparatus.
The feed water supply means may comprise a feed pump and associated metering means for providing feed water at the desired rate. Preferably the supply includes a feed water pump means driven by said positive displacement motor.
The output of the feed water pump means may be increased by increasing pump speed, increasing the stroke of the pump or by providing a stand-by pump.
Conveniently the feed water pump means includes a primary feed water pump arranged to continuously supply feed water to the steam generator whilst in operation and a secondary feed water pump arranged to intermittently supply feed water to the generator under control of said superheater thermostat.
The superheater thermostat is arranged to actuate said injection circuit valve means to allow fluid flow therein and to simultaneously actuate said secondary feed water pump to supply additional feed water to the steam generator coil. The injection circuit is arranged to by-pass a section of the coil, and preferably the inlet of the circuit is connected into the feed water heating zone of the coil and the outlet of the circuit is connected into a fast moving steam zone. The feed pumps are driven by a proportional drive so that good control can be obtained with the injection point located as far back as at a point in the evaporator zone of the steam generator, despite the difficult conditions encountered in an automotive steam power system having widely varying power requirements. It is preferable that the amount of feed water in the injection circuit is limited so as not to deplete the amount of water upstream of the injection outlet, thereby assisting in preventing the production of superheated steam upstream of the injection outlet.
The present invention allows close control over:
i. the amount of fluid injected by the injection circuit, and
ii. the amount of additional feed water administered by the secondary feed water pump thereby giving a rapid response to shortage of water signalled by the outlet thermostat in the steam generator. Furthermore, the additional feed water acts as a follow up to the rapid response provided by the injector.
It has been found that this injector-outlet thermostat system does give a rapid response to shortage of feed water, however fluctuations in the final steam temperature may still occur especially where the coil is composed of lightweight tubing having little heat reserve. To reduce these fluctuations even further, there is also provided by the present invention means for more accurately controlling the proportioning drive motor over its speed range by compensating for the effects of steam leakage and the effects of back pressures in the exhaust steam line to the drive motor at low and high motor speeds respectively. Said means includes a bypass valve or an electric/motor/generator controlling the speed of the motor over the middle or middle and high speed range. Conveniently the drive motor is a simple rotary motor.
The invention will now be described in greater detail having reference to the accompanying drawings.
FIG. 1 is a semi-schematic view of an overall steam plant showing various auxiliaries arranged in accordance with the present invention.
FIG. 2 shows a steam generator temperature curve of steam temperature against heat added to the steam generator.
FIG. 3 shows a feed water pumps performance curve of water pumped Q as a percentage of total weight of steam required R against power.
FIG. 4 also shows a similar curve of water pumped Q as a percentage of total weight of steam required R against power showing the effect of bypass correction by a metering valve of the feed pump drive motor.
FIG. 5 shows a curve similar to that in FIGS. 3 and 4 in which speed correction in the middle and high speed range of the feed pump drive is obtained by an electrical generator.
FIG. 6 shows a sectional view of a metering valve calibrated to relieve high back pressures in the exhaust steam line and also bypass some exhaust steam to provide the correction shown in the curve depicted by FIG. 4.
Referring to FIG. 1 a steam generator 10, is adapted to supply steam to an engine 11, by means of pipe 12 and throttle valve 13. A feed water system comprising a tank 19, a positive displacement pump 16, and feed water pipe line 17, is adapted to feed lower tubes 18 of the steam generator through pre-heaters 20 and 20a disposed in the exhaust conduit 21 leading from the engine 11 and engine 24 respectively.
The feed water pump arrangement 16 shown in FIG. 1 includes an auxiliary or secondary feed water pump 40 in parallel with a primary pump 16a, both pumps being preferably driven by the common drive 26. In this arrangement the secondary pump will run free while solenoid 43 is energized. Solenoid 43 is arranged to actuate an armature 42 constituting a valve controlling inflow of water to the pump 40 from tank 19. The current to the solenoid 43 is controlled by superheater thermostat 14. Thermostats 14 and 27 are connected to pivotable arms 15 and 28 respectively arranged to actuate contactors 14b and 27b between two way contact points 14a and 27a. The solenoid 43 is connected to the power supply through contactor 14b and is energised whilst the superheater thermostat 14 is sensing a temperature lower than a preset maximum.
If the preset maximum temperature is exceeded the arm 15 moves a sufficient distance to open the solenoid circuit and practically simultaneously close the injector valve circuit through the other contact point 14a. The injector valve, 52 which is also preferably solenoid actuated as shown at 50, is energized, provided the safety thermostat contactor and contact 27a and b are in the normal position as shown. Safety thermostat 27 is arranged to sense overheating in that part of the steam generator coil which is connected in parallel with the injection line. Alternatively, instead of closing the injector line the water injection point may be temporarily varied to a position upstream of the normal point (not shown). Alternatively the safety thermostat 27 may be arranged to reduce or cut off (not shown) the output of the burner 31. Thus, if the safety thermostat 27 senses a temperature above a predetermined maximum the solenoid circuit 50 is opened by movement of contactor 27b away from contact 27a thereby opening the circuit to the injector solenoid 50 and causing the injector valve 52 to close injector line 51 and thus restore full feed to the by-passed section of the coil.
A manually controlled switch 58 is provided to control operation of the system. An outlet steam pressure switch 53 is provided which is arranged to open and close switch 54 and disconnect and reconnect the burner motor with the power supply, when steam pressure exceeds a predetermined maximum, or falls below a predetermined minimum respectively.
As a precaution the superheater themostat 14 is also arranged to control operation of the burner motor. This control is shown in FIG. 1 comprising a burner switch 55 controlled by arm 15 connected to superheater thermostat 14. The contacts of the burner switch are opened upon the superheater thermostat 14 sensing a steam temperature in excess of (by a predetermined amount) the temperature of the steam which causes the superheater thermostat 14 to open contact 14a. Thus, the opening of the burner switch 55 is a second stage operation which shuts off the burner 31 as a back up to the pressure switch control 53 and the introduction of injector water and additional feed water if (despite the introduction of additional water) the steam temperature continues to rise to an undesirable level.
The exhaust conduit 21 carries exhaust steam from the engine 11 to rotary motor 24 through heat exchanger 20a and thence to a condenser 23. Water from the condenser 23 is returned to feed water tank 19. The rotary motor 24 drives a shaft 25 which in turn drives feed water pump 16 through belt 26. The rotary motor is arranged to drive other auxiliaries such as condenser fan 33, motor generator 56 and the like. The condenser fan drive may come from either side of the one way clutch 57.
The motor generator 56 operates as a motor at starting primarily for driving the feed water pump 16. It may operate as a generator for charging the battery power supply during normal running of the system and also may be used for a further useful purpose in governing the speed of the rotary motor. This latter purpose will be described in greater detail later. The one way clutch 57 is provided to transmit drive from the rotary motor to the motor/generator 56 and feed water pump 16 and condenser fan 33 when the rotary motor 24 is producing power but will not transmit drive from the motor/generator 56 when operating as a motor, as at start thus avoiding unnecessary load on the motor/generator 56. The one way clutch 57a is arranged to free wheel and thus prevent the burner motor 36 from driving the auxilaries on the other side of the clutch 57a.
The present invention has analysed the operation of the various components of the above described system in providing a steam generator control system consisting of:
1. a burner preferably of the ON/OFF type primarily controlled by a device responsive to generator steam pressure, and also an overriding temperature controller responsive to steam temperature in the superheater zone.
2. an exhaust steam rotary motor system preferably driving two feed pumps.
Operation of the feed pump has already been described in which one pump 16a is operable to pump water whenever it is rotated whilst operation of the second, auxiliary, pump 40 is under the control of the superheater thermostat 14 in the generator coil 18. The arrangement is such that the superheater thermostat 14 also controls the flow of injection water through the parallel injector circuit 51 on the generator coil at the same time as the second feed water pump 40 is operating dependant upon normal temperature conditions in the bypassed section of the generator coil.
The analysis of the variable components controlled by the invention is best shown by reference to various equations as hereinafter described in which the following symbols will be used.
P1 = rate of water by weight pumped by first pump.
P2 = rate of water by weight pumped by second pump.
Ew (extra Water) = rate of water by weight injected through injector nozzle when water is flowing through injector circuit.
R (Requirements) = rate of steam by weight passing out of the steam generator.
I. The first consideration to be outlined here is the amount of extra water EW to be injected as compared with the water pumped by the second pump P2. That is the ratio of EW to P2.
It has already been mentioned above that one reason for the long delay in response of the superheater thermostat situated in the superheater zone to the change in feed water quantity into the base of the steam generator is due to "compressibility" of the fluid between these two points.
To illustrate a point, it could be said that the effect of a change in "base" water feed is similar (in that part of the steam generator containing compressible fluid, i.e., steam) to that of a wave front carrying a higher level (high tide) or a lower level (low tide) of the density of the fluid behind the wave front.
The wave can be considered to be traveling at the speed of the actual fluid through the steam generator. In the evaporative steam zone where the dryness fraction of the steam is low, the velocity will also be low.
The response of a thermostat situated in the superheater zone to a change in feed from a water injector located at a point after which the steam is of a dryness fraction of 50 percent or more, or superheated, is rapid. In this case, the steam speed is relatively high and a short time period only is required before mixture is carried from the injector point to the thermostat. With such a rapid response, the thermostat control may turn water injection on and off rapidly enough such that there will be no resultant large fluctuations in the final steam temperature.
It has been found that an additional quantity of water, only up to a rate approximately equal to that fed by the water injector can be fed into the base of the steam generator in step with water injection fed directly from a feed water supply, by dividing the additional water fed into the base. Alternatively all additional water is fed into the base and injection water is obtained from a feed water heating economiser zone as shown in FIGS. 1 and 2. In either case the following discussion generally follows although it particularly applies to the first case in which the additional base water is divided.
The superheater thermostat 14 acts to regulate the quantity of injection water required and could be said to act as an "early warning" regulator on the amount of water entering the base of the steam generator. If, on the other hand, the increase in the amount of water fed into the base of the steam generator is greater than the injection water quantity, there is the likelihood that, when this increased flow "comes throug" to the superheater zone thermostat, it will be too much and it will be too late to turn it off soon enough to prevent an excessive down swing in final steam temperature following.
Thus the water control system operates as follows: P1 is always less than R and, from the above, the additional water feed into the base of the steam generator when the second pump is pumping, -- i.e. (P2 - EW) must be equal or less than EW i.e.: P2 - EW .ltoreq. EW and EW .gtoreq. .50 P2 (1)
ii. of major importance in the control of a steam generator is the ability of the system to control events following a change from the pumping of a smaller quantity of water (such as P1) to a greater quantity of water (such as P1 plus P2). Consider the case where temperature is rising at the superheater outlet thermostat and the latter has caused the second pump and the injection water to be switched on. The flow of steam after the injection point in the steam generator must match "R" without waiting for additional feed from the base of the steam generator. The worst case would be where the flow in the steam generator just before the water injection point may have fallen to "P1 " (low tide).
To satisfy the above, P1 + EW.gtoreq.R. If this requirement is not met, in the above case, temperatures will continue to rise and the thermostat override control will shut off the burner. This will lead to a loss of available power if the steam generator pressure has fallen into the range where burner operation is otherwise required.
In order to allow for such factors as steam generator thermal delay, P1 + EW should have some margin over R, especially if a more rapidly fluctuating injection water control is required in order to assist in smoothing out fluctuations of water feed through the base of the steam generator. With 10 percent margin, P1 + EW.gtoreq.1.10 R.
In addition to the above a further factor must be considered in the case where an ON/OFF burner is used. Consider the case where the system is running at light load, the burner is operating and the second pump and injection circuit have been switched on by rising temperatures in the superheater thermostat 14. In the evaporative zone, a temperature change of from, say, 544 DEGF to 587 DEGF, i.e. 43 DEGF, is required to raise boiling point pressures from 1000 psi. to 1400 psi. at which later pressure it is assumed the burner would be switched off. The above temperature rise may be achieved with a corresponding temperature rise at the superheater thermostat of twice this amount i.e. 86 DEGF. (depending on steam generator tubing layout etc.). Now it is not desirable to have to set the temperature for operation of the burner over-ride control at a large amount above that temperature at which the control operates the second pump and water injector, in order that the burner over-ride will not operate under normal conditions.
Sufficient pressure rise throughout the steam generator can be obtained with a more moderate temperature rise at the superheater thermostat 14, if the quantity of water and steam in the steam generator is increased. Thus, if P1 + EW is increased to be greater than R by an additional margin, (i.e. -- feed will tend somewhat to better match the momentary burner rate rather than the steam output rate) -- satisfactory results may be achieved with closer temperature settings for the pump/injector control and the burner over-ride control.
Thus allowing the further margin for the ON/OFF burner system,
P1 + EW.gtoreq.1.20 R (2)
considering the above case but with a modulating burner which matches the load more closely, temperatures would not be expected to rise significantly with the two pumps and EW feeding with the relation P1 + EW.gtoreq.1.10 R. Thus it would be expected that satisfactory results would be achieved with the quantity P1 + EW less than for the case with an ON/OFF burner system. For an ON/OFF burner system at full load, in which water and burner rates are more closely matched, a relation similar to that applying to the modulating burner system would see applicable.
It should be noted that there are many factors which have some effect in connection with the above relation (2). The applicant has found, however, that experimental results do tend to support the above reasoning.
III. To avoid internal steam generator temperature peaks, control should be exercised over the proportion of injection water provided.
The following method calculates the maximum rate of injection water "EW" injected so that the dryness fraction qB of steam just before the injection circuit outlet is 100 percent i.e. just not superheated.
Having reference to FIG. 2 the full line "I" in the graph indicates water and steam conditions throughout the steam generator heating surface under steady conditions when all feed water is delivered into the bottom of the steam generator, and is equivalent to the burner evaporation capacity at the particular load. Note that a burner controlled on an "ON/OFF" basis can give roughly similar results in matching the load as a modulating burner. The dotted line "II" shows the variation from the above when total feed water pumped equals the burner capacity as before but part of the water "EW" is taken from a feed water heating zone (as is good practice for injection systems) at point "W" and injected into a point "Z" immediately after which the dryness fraction is qA = 60 percent. Steady conditions are again assumed.
The percentage of heat received by water-steam following line I between points W where temp. = 450 DEGF and Z = 11.2 + 28.8 = 40 percent, producing steam at q = 60 percent at Z.
Considering unit weight of water/steam, the percentage of heat to produce steam at q = 100 percent from water at 450 DEGF = 11.2 + 48 = 59.2 percent of total heat added.
It can be seen that, if heat supplied to the steam generator section between W and Z remains constant, and quantity of water passing along this section drops in the ratio of 40 to 59.2 i.e., drops to 40 + 59.2 = 67.6 percent of its former value, steam of qB = 100 percent will be formed just before A i.e., EW = 100 - 67.6 = 32.4 percent of total water pumped. If EW>32.4 percent, steam will superheat just before Z.
It is possible to use an "earlier" injection point to enable EW to be greater. However, greater delay in response to the thermostat would occur. Conversely, with a "later" injection point, EW would have to be less but response delay would be less. It may be desirable to reduce EW to conform with the considerations discussed in paragraph I above. It is considered that the injection point shown is approximately at the optimum position.
It could be argued that a small superheat before Z could be tolerated. Care is needed if this is assumed for the design based on an "ideal" graph. The above examples assumes steady conditions and, in practice conditions are not steady. Variations can occur such as load changes which because of factors such as inertia in flow response to change of load, can lead to effects causing steam before Z to become wetter or drier (superheated) than estimated for steady conditions. A margin of safety is required over the "ideal" graphs shown for steady conditions. Thus, from the above considerations, it appears that the water injector control could control 2 .times. 32.4 = 64.8 percent only of the total feed water. An additional control system is therefore needed.
From the above calculations, it can be seen that, to avoid internal temperature peaks, base water f .gtoreq. .676 R. Since base water feed may, at times, approach P1 (low tide) thus P1 .gtoreq..676 R.
Under some conditions, P1 + EW may be approximately equal to R, then P1 .gtoreq..676 (P1 + EW) from which
EW.ltoreq..48P1 (3)
under conditions such as may occur immediately after start-up, the flow reaching "Z" on the curve shown in FIG. 2 from the base of the steam generator, may temporarily be <P1. The temperature before "Z" would be expected then to rise and the safety thermostat 27 (FIG. 1) would possibly operate.
IV. Considerations involving reductions in steam temperature fluctuations
Some causes of temperature fluctuations in the steam leaving the steam generator are: (a) Response-delay in the superheater thermostat 14 in sensing the correctness of the mixture at "Z," and (b). The magnitude of the "error" in the mixture reaching the superheater thermostat 14.
Assuming a fixed response-delay time, reductions in temperature fluctuations can be achieved by bringing P1 closer to R and minimising EW and P2. Thus there is argument for P2 to be less than P1 i.e. -- Pumps of different capacities, referred to in more detail later.
V. Consideration of the quantities and relationships between P1 P2 and R as effected by Rotary Motor Characteristics
The graph (FIG. 3) shows the effects of leakage and back pressure on the rotary/motor/feed pump/condenser-fan drive system. The effect of leakage is large at the low powers thus leading to low rotary motor speeds. The high back pressure of the fan, rising as the square of the speed, causes a rapid increase in back pressure required to operate the rotary motor at high powers again leading to reduced rotary motor speeds.
It can be seen from the graph, and using the simplified considerations the useful range is that in which P1 <R and P1 + P2 >R, it can be seen that difficulties in obtaining a useful wide range increase as P2 becomes small in proportion to P1. (See later for rotary speed correction devices which assist in overcoming this factor).
The above considerations, I to V are in themselves narrow ones. Account is not taken of such factors as failure of one pump, thermal storage in the steam generator tubes, changes of steam zone positions with changes of load, inertia of the steam generator contents in following load changes. Because of the changes in rotary motor system performance with load, P1 will not bear a fixed relation with R, for example.
The steam generator system described in this specification, however, is protected by the action of a "safety" thermostat and the superheater thermostat as well as a steam pressure switch as previously described. Rapid accommodation to load changes is made with the rapid action of the water injection control system.
Summarising the relations evolved above:
EW.gtoreq..50 P2 (1)
p1 + ew.gtoreq.1.2 r (2)
ew.ltoreq..48 p1 (3)
using a system with the position of the water injection point "Z" as shown in FIG. 2, (i.e. -- so that the dryness fraction after "Z" is 0.60 with feed of water matching output for steady conditions,) and using twin feedpumps so that P1 = P2, with EW = .5 P2, relation (1) will be satisfied and relation "3" will be approximately satisfied. From relation "2" --
1.5 P1 .gtoreq.1.2 R and P1 .gtoreq..80 R
Note that, if EW increased, relation "3" is not satisfied. This means that there is a possibility, under abnormal conditions, of a temperature peak before "Z." The safety thermostat would operate if necessary but this may cause a more serious loss of good control than if EW was not increased. In the latter case, the superheater thermostat may reduce burner output if required under abnormal conditions.
Some more latitude can be allowed for EW in a system using pumps of different sizes. With P1 = 1.15 P2, from relation "3,"
Ew.ltoreq..48 p1
.gtoreq..552 p2
thus EW may be from 0.50 to 0.552 P2.
For EW = 0.50 P2, from relation "2", P1 .gtoreq.0.837 R,
For EW = 0.552 P2, from relation "2," P1 .gtoreq.0.81 R.
EARLIER INJECTION, AND MULTIPLE INJECTION
With water injection earlier than shown, (FIG. 2) EW can be safely increased and a larger margin of operation of P1 as a function of R can be achieved. Response delay can be reduced by injecting through more than one injection point.
EXAMPLE
First Injection Point such that, under steady conditions, with no water injection, dryness fraction of steam = 0.50. Using a method similar to that used for finding relation "3," total EW.ltoreq..66 P1. Half of EW can be injected through a second injection point after which, under steady conditions, dryness fraction of the steam would be, say 0.75.
METERING AND PROPORTIONING OF INJECTION WATER
The injection water line 51 in FIG. 1 incorporates a metering jet which, in the preferred arrangement is the orifice of the solenoid control valve 52 see FIG. 1. This jet is designed to allow the passage of quantities of water equal to approximately 0.50 P2 or as calculated by the use of the above relations.
The method of estimation of the jet size may be as follows:
A percentage load is assumed and the corresponding pressure drop of the water and steam passing through the steam generator proper, between W and Z, FIG. 2. is calculated. The jet size is then calculated so as to pass the correct amount of water at the estimated pressure drop.
EFFECT OF LOAD CHANGE ON EW
The pressure drop from W to Z will vary approximately as the square of the load. The weights of water and steam passing through the steam generator proper between W and Z and also through the water injector will vary but will remain approximately in the same proportions.
EFFECT OF PRESSURE DROP ON EW
At low steam generator pressures, such as may occur immediately after start-up, pressure-drops through the steam generator will be higher (for the same load) due to the lower density of the steam and the higher steam speeds. The proportion of water through the water injector will thus tend to rise. However, the action of the safety thermostat will protect the steam generator if there is any significant upward surge of temperature because of the above.
Referring to FIGS. 4 to 6, FIGS. 4 and 5 show curves indicating the effect of speed correction of the rotary motor 24 (see FIG. 1) in the middle of the range where the speed of the rotary motor tends to be higher than required for proportional control of the feed water pump 16 compared with the steam requirement of the generator. FIG. 4 shows by the dotted line, correction by a bypass or leak valve which has the effect of causing the speed of the rotary motor 24 to remain more closely proportional to the steam requirement, substantially over the useful load range of the power unit.
FIG. 5 shows speed correction by the connection of motor/generator 56 (see FIG. 1) into the rotary motor drive circuit. The motor/generator 56 when operating as a motor is controlled automatically so as to cause feed water to be pumped into the steam generator at a rate approximately equal to 20 percent of the full load rate at such times as the steam generator pressure is substantially below normal and the steam temperatures are above normal. These conditions may occur just after initial start up.
The generator of the motor/generator is operative to impose a torque load on the rotary motor in the middle speed range which is inherently reduced because of the lower torque demand of the generator at higher rotary motor speeds.
The generator may be of the third brush or constant current type and "cut in" of the generator at low speeds may be suitably delayed to reduce torque load on the rotary motor.
FIG. 6 shows a metering valve for positioning in the exhaust steam circuit in parallel to the rotary motor. The valve includes a chamber 60 having a piston 61 therein, the piston 61 is movable between two positions under the controlling influence of biasing springs 62, 63 and steam pressure. The chamber is ported at 64 to allow leakage of steam past the piston 61 at a predetermined pressure in the exhaust steam circuit representing the middle speed range of the rotary motor, thereby bypassing the rotary motor with some of the exhaust steam. The position shown in FIG. 6 is an intermediate position.
With back pressure higher than those normally encountered at full load, such as short term exhaust pressure surges, the piston may take up an extreme position thereby by-passing a considerable amount of steam and relieving the pressure surge.
GB1282613
IMPROVEMENTS RELATING TO THE CONTROL OF EXPANSION RATIO IN ROTARY MOTORS
IMPROVEMENTS RELATING TO THE CONTROL OF EXPANSION RATIO IN ROTARY MOTORS
A motor driven by a compressible fluid such as steam or air has a rotary structure 4 attached to an output shaft 21 and furnished with sealing segments 4a that engage with blades 3 supported by a fixed shaft 2. A proportion of the fluid in a supply line 6a is diverted through a flow-restricting passage 8 and a passage 9 into one of the interblade chambers 5 between an inlet port 6 and an outlet port 7. When it is necessary to augment the torque at the shaft 21 e.g. on starting, the flow through the passage 9 is increased by a valve 11 opening to connect a passage 10 to a passage 15, this being due to a greater pressure-differential occurring between the line 6a and a passage 13 and consequent deflection of a spring-loaded diaphragm 12. The effect produced by the increased flow through the passage 9 is comparable to that of "late cut-off" in a conventional reciprocating steam-engine. Alternatively, the valve may be such that it opens in response to a rise in the pressure in the line 6a, (For Figures see next page)
This invention relates to rotary motors ob the positive displacement type using steam or compressed air or other expansible gas or vapour as distinct from rotary turbines, in which motors the expansion ratio can be varied without the aid of intake valves such as rotary inlet valves or inlet valves operated by means of link motions or camshafts.
In a rotary motor or reciprocating engine of the positive displacement type which is used over a wide range of speeds as, for example, from zero to the maximum design speed, it is desirable to use a small expansion ratio or "late cut-off" of the high pressure working fluid in order to obtain more positive starting, high overload torque and better smoothness on starting. A smaller expansion ratio or "later cut-off" is also desirable, apart from when starting, in order to run against loads heavier than normal. Under normal loads, it is desirable to run on a higher expansion ratio (early cut-off) so as to obtain more economical use of the working fluids.
In some rotary and reciprocating engines, changes of cut-offs arc obtained, for example, simply by varying the aiTangements of link motions or by changing the positions of cams or, on engines fitted with rotary valves, by changing the position of rotary valve sleeves.
However it is advantageous if intake valves can be eliminated and to rely solely upon porting of the rotary engine thereby reducing initial and ensuing maintenance costs.
This invention has for its principal objective to provide a rotary motor of the type specified in which the expansion ratio can be varied in accordance with load and speed requirements.
With the principal objective in view there is provided according to the present invention a rotary motor of the positive displacement type driven by expansion of a compressible working fluid introduced from an inlet supply line under pressure through an inlet port of the motor into a first expansion stage, said motor having at least one additional expansion stage between said first expansion stage and an outlet port of the motor spaced from said inlet port, the improvements comprising a bleed passage from said inlet supply line leading working fluid to at least one said additional expansion stage in the motor, said bleed passage including means for automatically controlling the amount of working fluid supplied therethrough such that during periods of low speed and/or high load the expansion ratio is reduced by increasing the amount of working fluid admitted to the motor.
Conveniently a pressure sensitive valve is provided in said bleed passage said valve being influenced by the load on the motor.
The bleed passage may be restricted and said restriction may be provided in parallel with said valve to provide a continuous flow of working fluid which is effective at low motor speeds to substantially increase the amount of working fluid admitted to the motor.
Said pressure sensitive valve is convenientLy sensitive to working fluid inlet or expansion stage pressure, or a pressure differential between inlet and expansion stage inlet pressures by which effects of later cut-off are obtained upon increase in load.
It will be understood that according to the invention introduction of the working fluid may take place at several succeeding stages of expansion to a lower pressure in the motor.
A practical embodiment of the invention now to be described is a blade type rotary positive-displacement motor driven by steam or air, however it will be understood that the invention can be applied to various kinds of rotary positive displacement motors in which there is at least two stages of expansion of the working fluid. The embodiment is described having reference to the accompanying diagrammatic drawings.
Figure 1 is an end sectional view on line I-I.
Figure 2 is a sectional elevation taken on line 11-11.
There is provided a stationary cylindrical ported chamber 1. In sealing contact with said chamber 1 a plurality of blades 3 are provided which extend radially from bearing bosses 3b mounted on shaft 2 above centre 0 enabling said blades 3 to move about centre 0 independently of each other in the direction of rotation shown. The blades 3 are constructed with their inner portions and bosses 3b forked to fit inside each other along the shaft 2. The blades 3 pass through sealing segments 4a mounted in a rotatable cylindrical structure 4 which is integrally formed with an end flange 4c at the drive end and with a removable flange 4d at the opposite end. The blades 3 are stepped dozen at 3a in overall width (taken along the axis) to clear inner portions 4b of the end flanges of the structure 4.At the driving end of the structure 4 an output shaft 21 on centre C extends from the end flange 4c.
Shaft 2 is held into end wall 23 of chamber 1 by means of a nut 22. Output shaft 21 runs in bearings machined into an extension 24 of the opposite end wall of chamber 1.
The ports, inlet 6 and outlet 7 in the chamber 1 are placed approximately opposite one another.
It is preferred that a sufficient number of radial blades 3 are provided so that at least two expansion stages are formed between the inlet port 6 and the exhaust port 7.
Accordingly smooth running of the motor is obtained despite the lack of intake and exhaust valves as well as providing a greater expansion ratio.
The bleed fluid passes through a restricting passage 8 of a definite predetermined minimum control area in bleed line 9 leading from the main supply line 6a to a later expansion stage. The amount of bleed fluid is increased by opening up of additional passage area controlled by means of a pressure sensitive valve
11 which may be diaphragm controlled and inter-connects passage 10 to passages 9 and 15 through needle valve member 14. The diaphragm 12 may be sensitive to the pressure existing at a datum point such as the fluid intake (not shown) or to a pressure differential such as the difference between the pressures at the intake 6 and at the point of admission
16 of the working fluid as shown in the embodiment illustrated.In this embodiment the diaphragm 12 is subjected to the pressure in passage 10 on the one side and the pressure in passage 13 on the other side. Since the pressure of working fluid bears a direct relationship to the load on the motor, the valve 11 is in fluenced by the load on the motor. The diaphragm is adapted to move in response to the bias created by the pressure differential whereby the needle valve member 14 is moved to open or close passage 15. If necessary a compression spring 17 may be provided to ensure that the needle valve member 14 is closed at the appropriate time.
A major controlling factor over the maximum expansion ratio on the type of rotary motor illustrated is the number of blades 3.
The larger the number of blades, the greater the expansion ratio obtainable.
It will be appreciated that the bleed fluid flowing in the restricted passage S, will have little effect at normal running speed, the volume of flow being of but a small proportion of the total flow of working fluid through the motor. Thus the bleed passage is mainly effective at low speeds as desired.