EP0085800B1 - Einheit aus Freikolbenmotor und Pumpe - Google Patents

Einheit aus Freikolbenmotor und Pumpe Download PDF

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Publication number
EP0085800B1
EP0085800B1 EP19820300655 EP82300655A EP0085800B1 EP 0085800 B1 EP0085800 B1 EP 0085800B1 EP 19820300655 EP19820300655 EP 19820300655 EP 82300655 A EP82300655 A EP 82300655A EP 0085800 B1 EP0085800 B1 EP 0085800B1
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EP
European Patent Office
Prior art keywords
engine
pump
piston
fluid
compression
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Expired
Application number
EP19820300655
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English (en)
French (fr)
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EP0085800A2 (de
EP0085800A3 (en
Inventor
Robert D. Vanderlaan
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Pneumo Abex Corp
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Pneumo Abex Corp
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Priority to EP19820300655 priority Critical patent/EP0085800B1/de
Priority to DE8282300655T priority patent/DE3276338D1/de
Publication of EP0085800A2 publication Critical patent/EP0085800A2/de
Publication of EP0085800A3 publication Critical patent/EP0085800A3/en
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Publication of EP0085800B1 publication Critical patent/EP0085800B1/de
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B17/00—Pumps characterised by combination with, or adaptation to, specific driving engines or motors
    • F04B17/05—Pumps characterised by combination with, or adaptation to, specific driving engines or motors driven by internal-combustion engines

Definitions

  • the present invention relates generally to a hybrid power system for generating pressurized hydraulic power and, more particularly, to free piston engine pumps in which energy of combustion in a power cylinder is converted into hydraulic energy.
  • a free piston engine pump (hereinafter abbreviated FPEP) the motion of the engine piston(s) is at least substantially directly delivered to hydraulic pumping elements, usually, without crankshaft and connecting rod arrangements of conventional rotary engines.
  • the hydraulic power developed may be used for vehicle propulsion and auxiliary equipment operation as well as for other purposes.
  • US-A-4 307 999 It is generally known from US-A-4 307 999 to provide a free piston engine pump including a primary piston movable in pump chambers and valve means (13, 15, 17) for controlling the flow of fluid into and out of the pump chambers, said valve means comprising first check valve means (17) for passing inlet flow of relatively low pressure fluid to the first pump chamber (4i), and second check valve means (15) for passing outlet flow of relatively high pressure fluid from the first pump chamber.
  • the valve means (13) of US-A-4 307 999 generally corresponds to the valve means (69) and check valve means (65) of the present invention in combination.
  • the present invention is concerned with optimizing the efficiency of a FPEP and providing versatility and facility of operation and use thereof.
  • the FPEP of the present invention includes an engine for producing mechanical work during a power stroke and a pump responsive to the engine work for pumping fluid during the power stroke.
  • the intake ports and exhaust ports of the engine combustion chamber are at opposite ends thereof resulting in unidirectional or uniflow scavenging of the engine cylinder.
  • Valving controls hydraulic input and output paths of the pump to permit selective operation in a primary high flow and a secondary high pressure mode of operation, preferably while maintaining substantially constant the product of output pressure and flow; the valving also may be employed selectively to control cycle rate, i.e. the number of cycles per unit time, starting either in the primary mode or secondary mode, intermittent cycling, and compression energy boost. Pumping may be effected during the entire power stroke and in the normal operating region compression energy is supplied during the entire compression stroke.
  • an in line opposed piston free piston engine pump system comprising a free piston engine including an engine cylinder having a linear axis, a pair of engine pistons movable in said engine cylinder along such axis toward each other during a compression stroke and away from each other during a power stroke; a separate pump means associated with each engine piston for pumping fluid, each pump means including a pump cylinder axially aligned with said engine cylinder, a pump piston coupled to a respective engine piston for movement therewith and forming first and second pump chambers within said pump cylinder, valve means being provided for controlling flow of fluid into and out of said pump chambers, said valve means comprising first check valve means for passing inlet flow of relatively low pressure fluid to said first pump chamber, and second check valve means for passing outlet flow of relatively high pressure fluid from said first pump chamber, characterized in that a first selectively operable valve means is arranged in parallel with said second check valve means for selectively bypassing the latter, third and fourth check valve means respectively for passing inlet flow of relatively low pressure fluid
  • a deformable bladder-type compression accumulator may be used for storing energy during the power stroke and returning the same for compression; such accumulator contains a compressible fluid the pressure of which is con- trollably adjustable to control compression energy.
  • Total control of the energy put into compression to establish compression ratio and the related pressure and temperature condition in the cylinder enable optimization of engine efficiency minimizing compression losses and controlling operating pressure profile in the engine cylinder; moreover, the rate at which compression energy is applied may be controlled to establish the velocity and acceleration profiles of the engine pistons during compression stroke enabling cycle rate variability.
  • aspects of the invention include synchronizing the pistons of an opposed piston type of FPEP preferably without ordinarily substantially loading the synchronization apparatus; acceleration boost of the pistons at the start of a compression stroke; an energy absorber for excessive energy during an abnormal power stroke; and a reset valve and actuator arrangement for a FPEP. Still additional aspects relate to control features whereby a plurality of engine and/or pump parameters may be monitored electronically and operation accordingly electronically controlled and to the interfacing or pairing of plural FPEP's resulting in reduced pressure pulsations, versatility allowing less than all of the FPEP's to operate at a given time, and general efficiency by combining elements and functions.
  • a FPEP in accordance with the invention is generally illustrated at 1.
  • the FPEP 1 has an engine portion 2 and hydraulic pump portion 3.
  • the engine portion 2 includes an engine cylinder 4 in which a pair of engine pistons 5, 6 move linearly or axially and between which a combustion chamber 7 is formed.
  • a fuel injector 10 injects fuel into the combustion chamber 7.
  • Air intake ports 11 at the righthand end of the combustion chamber 7 provide passage for air into the same, and exhaust ports 12 at the opposite end of the combustion chamber 7 permit exhaust gases to exit via an exhaust line 13. With the intake and exhaust ports 11, 12 located at opposite ends of the combustion chamber 7, uniflow or unidirectional scavenging is achieved.
  • the engine portion 2 is of the opposed piston type, whereby during a compression stroke the engine pistons 5, 6 are urged toward each other in the cylinder 4 reducing the size of the combustion chamber 7 and, therefore, increasing the pressure and temperature therein to effect compression ignition of the fuel injected by the fuel injector 10, thereby to initiate a power stroke.
  • the engine pistons 5, 6 are driven by the energy of combustion oppositely axially in the cylinder 4.
  • the exhaust products will exit the combustion chamber 7 via the exhaust line 13, and when the engine piston 5 subsequently opens the intake ports 11, air will enter the combustion chamber 7 to effect the desired scavenging after which the next compression stroke usually will commence.
  • the engine pistons 5, 6 are of equal mass and those parts of the engine portion '2 movable with the engine piston 5 are of a mass equal to those parts movable with the engine piston 6 thereby effectively naturally to mass balance the engine portion 2 about a centerline 14, which is perpendicular to the linear axis 15 of the engine portion.
  • synchronizers 16 mechanically interconnecting the engine pistons 5, 6.
  • Each synchronizer 16 includes a pair of racks 17, 18 connected to the engine piston 6 for linear movement therewith and a pair of pinion gears 19, 20 for rotation by the respective racks 17, 18.
  • the pinion gears 19, 20 are coupled by a shaft 21, which rotates with the pinion gears and is in turn coupled via a pair of bevel gears 22, 23 to turn a coupling shaft 24.
  • the coupling shaft 24 in turn is connected to a similar arrangement of racks, pinion gears, shaft and bevel gears (not shown) like those identified by the reference numerals 17-23 associated with the engine piston 5.
  • Compression energy during a compression stroke is preferably applied by both engine pistons 5, 6 independently of the synchronizer 16, and during a power stroke the energy of combustion directly urges both engine pistons 5, 6 relatively outwardly in the cylinder 4.
  • the synchronizer 16 desirably effects its synchronizing operation generally maintaining a balanced uniform operation and movement of the engine pistons 5, 6 usually without any appreciable mechanical loading or forces on the various components of the synchronizer 16.
  • Air to support combustion passes through the air filter 25, is compressed by a compressor 26 and is delivered usually at several, preferably as much as three, atmospheres pressure via air line 27 and intake ports 11 into the combustion chamber 7.
  • An after cooler 28 in the air line 27 provides a cooling or heat exchange function, visa-vis the inlet air and the combustion chamber.
  • air entering the air filter 25 also may pass via air line 29 and one or more reed check valves 30 to the back side chamber 31, 32 of each engine piston 5, 6, being drawn there during a compression stroke and being pressurized during a power stroke and, accordingly, forced at the end of a power stroke through the intake ports 11 further supercharging operation of the engine portion 2 in cooperation with the pressurizing function of the compressor 26.
  • each cycle of operation of the engine portion 2, including a compression stroke and a power stroke preferably is substantially uniform to enable relative tuning of the exhaust system 38 for optimum utilization of the energy contained in the exhaust products of combustion.
  • the righthand half 3R of the hydraulic pump portion 3 is illustrated in Fig. 1 and will be described in detail below associated with the engine piston 5; the other half 3L of the hydraulic pump portion 3, schematically shown in Fig. 13, may be substantially identiacl to that described.
  • control of both halves of the hydraulic pump portion 3 preferably will be simultaneously parallel.
  • the inlet hydraulic fluid line 40 and the outlet hydraulic fluid line 41 associated with the righthand pump half 3R preferably would be coupled to a hydraulic system, not shown, in parallel fluid relation with the lefthand pump half 3L.
  • relatively low pressure inlet hydraulic fluid at pressure P is directed to the inlet hydraulic fluid line 40, and the hydraulic pump portion 3 pumps hydraulic fluid at relatively high pressure P. via the outlet hydraulic fluid line 41 for use in the external hydraulic system.
  • the same includes a pump piston 42 slidably movable in a pump cylinder 43 is sealed relation thereto using a single conventional sliding seal 44.
  • a rod or shaft 45 mechanically connects the pump piston 42 with the engine piston 5 for linear in-line reciprocation therewith.
  • a compression piston 46 slidable with respect to a compression cylinder 47.
  • a compression fluid flow line 48 extends between the compression cylinder 47 and the compression accumulator 49. It is the purpose of the compression piston 46 and compression accumulator 49 to store compression energy, i.e. energy required to effect a compression stroke, during a power stroke and subsequently to deliver such compression energy to the engine piston 5 to effect such compression stroke after a power stroke has been completed.
  • a deformable bladder-like member 50 contains a compressible fluid, such as an inert gas or other gas; the bladder 50 in turn is contained in a rigid accumulator housing 51.
  • hydraulic fluid in the compression fluid flow line 48 is pumped by the compression piston 46 into the accumulator housing 51 effecting a deforming of the bladder 50 to compress the gas therein thereby storing compression energy.
  • the compression energy stored in the compressed gas in the bladder 50 is delivered via the hydraulic fluid in flow line 48 urging the compression piston 46 and, thus, the engine piston 5 to move toward the left in a compression stroke.
  • the seal 52 provides the desired isolation in that during both a normal power stroke and the compression stroke in the secondary mode of operation the pressure in the high pressure output accumulator 53 and compression accumulator 49 are preferably approximately equal.
  • a differential pressure of P c (the pressure in the bladder 50) minus P, does exist across the seal 52, such differential pressure preferably will be relatively small and ordinarily certainly less than one containing ambient pressure as a term.
  • At the front or pressure side of the pump piston 42 is a first pump chamber 60, and at the back side of the pump piston 42 is a second pump chamber 61.
  • Inlet check valves 62, 63 supply low pressure P, fluid in the flow direction shown from the inlet hydraulic fluid line 40.
  • a low pressure accumu- latory 64 coupled to the inlet hydraulic fluid line 40 stores inlet hydraulic fluid for supply to the hydraulic pump portion 3 while also minimizing pressure pulsations of the inlet fluid.
  • Outlet check valves 65, 66 are coupled at chambers 60, 61 to the outlet hydraulic fluid line 41 and high pressure accumulator 53.
  • Three selectively operable control valves 67-69 are used to control the operation of the hydraulic pump portion 3, as will be described in greater detail below.
  • valves 67-69 preferably either are electrohydraulically or mechanically actuated and provide large passageways through which fluid may flow thereby to effect rapid operation to flow opening or closing and to avoid pressure losses; the preferred valve is a ball type in-line valve.
  • Conventional pressure control valves 70 are coupled to the fluid flow lines illustrated to relieve excess and otherwise to alter fluid pressure, if needed.
  • pumping preferably is accomplished during the entire power stroke and compression energy is applied preferably during the entire compression stroke.
  • the primary pumping element is the pump piston 42 with the minimum number of seals illustrated.
  • high pressure hydraulic fluid is pumped out on the power stroke and drawn in on the compression stroke by action of the arrangement of check valves illustrated.
  • Check valve closing occurs preferably only at the ends of the stroke where piston velocity decreases uniformly to zero.
  • the valves have a natural decreasing flow profile so that when piston motion stops the valve is immediately seated thereby eliminating a tendency for backflow leakage when piston motion reverses.
  • output and input flow rates are continuous throughout the strokes with no discontinuity associated with alternate techniques that change energy in discrete levels during the strokes.
  • Full stroke pumping also reduces the peak hydraulic flow rate passing through the check valves and various flow passages thereby reducing hydraulic losses.
  • the hydraulic pump portion 3 supplies energy for effecting a compression stroke to bring the engine pistons 5, 6 toward one another thereby to effect compression of fuel and air in the combustion chamber 7. Thereafter, the energy of combustion drives the engine pistons 5, 6 relatively outwardly to expand the combustion chamber 7 in a power stroke during which fluid is pumped by the hydraulic pump portion, as now will be described in detail.
  • the hydraulic pump portion 3 has two distinct modes of operation, namely a primary high flow rate mode, which normally is used and is the more efficient mode of operation, and a secondary high pressure mode, depending on the setting of the control valves 67, 68 and 69.
  • Figs. 2A and 2B operation of the FPEP 1 and particularly the hydraulic pump portion 3 in the primary high flow rate mode is illustrated.
  • the control valve 67 is open and the control valves 68, 69 are closed.
  • energy stored in the form of compressed gas in the compression accumulator 49 is delivered via the hydraulic fluid in the compression fluid flow line 48 to drive the compression piston portion 46 of the pump piston 42 and, thus, the engine piston 5 to the left relative to the illustration effecting compression in the combustion chamber 7.
  • hydraulic fluid enters the first pump chamber 60 via the inlet check valve 62 while a lesser amount of fluid exits from the second pump chamber 61 via the open control valve 67.
  • high pressure fluid is pumped by the pump piston 42 and exits the first pump chamber 60 via the outlet check valve 65 as relatively low pressure fluid returns to the second pump chamber 61 via the open control valve 67.
  • the inlet fluid provided via the inlet hydraulic fluid line 40 is desirably at relatively low pressure preferably stabilized by the low pressure accumulator 64 (Fig. 1), and the hydraulic fluid pumped from the first pump chamber 60 to the outlet hydraulic fluid line 41 will be at relatively higher pressure and may be used to do work in external equipment, not shown, or may be stored in the high pressure accumulator 53.
  • the control valve 67 is closed and the control valves 68 and 69 are open.
  • high pressure fluid enters the first pump chamber 60 through the open control valve 69 and high pressure hydraulic fluid also exits the second pump chamber 61 via the outlet check valve 66 and open control valve 68. Since the area of the pump piston 42 exposed in the first pump chamber 60 exceeds that exposed in the second pump chamber 61, the just-described flow of fluid will effect a net energy or work input during the compression stroke to supplement the compression energy provided by the compression accumulator 49, as was described above with reference to Fig. 2A. However, in the secondary mode, the pressure level P c in the compression accumulator 49 preferably would be reduced substantially in order to minimize losses, whereupon the principal compression energy is delivered from the outlet hydraulic fluid line 41.
  • the primary and secondary modes of operation may be compared assuming, for example, equal power input level for both modes of operation whereby the net hydraulic output work must be the same for both modes, disregarding losses.
  • the output work in each mode is proportional to the product of pressure times flow and input work equals output work. Therefore, if the volume of the first pump chamber 60 is twice that of the second pump chamber 61, the output pressure capability of the secondary mode of operation will be twice the pressure capability in the primary mode of operation.
  • a reset mechanism 75 associated with the pump half 3R the other pump half 3L also having a similar reset mechanism or connections to the one shown, is operated to position the engine pistons 5, 6 and pump pistons 42 outward, as is illustrated, for example, in Fig. 4.
  • Such outward positioning is accomplished by venting the fluid in the first pump chamber 60 to inlet pressure level P ; by a connection effected through hydraulic fluid line 76, chamber 77 of a selectively adjustable reset spool valve 78 and fluid line connection 79, which is connected to the low pressure accumulator 64, for example.
  • fluid pressure in the second pump chamber 61 is raised by supplying high pressure fluid from an external source (not shown) via fluid line connection 80, chamber 81 of the reset spool valve 78, reset actuator 82, check valve 83 and hydraulic fluid line 84, thereby providing adequate pressure to force the pistons outward against the resisting force of the compression accumulator 49.
  • the inlet P i , outlet P o , and compression accumulator P c pressure levels must first be established by conventional means which are not part of this disclosure.
  • the FPEP 1 is designed to start at minimum operating pressure level of, for example, 140 Bars (2000 psi) or greater. After the minimum pressure level for start-up has been established, a force level great enough to overcome the compression accumulator pressure acting overthe surface area of the compression piston 46 as well as all frictional forces is required to move the engine and pump pistons to the position illustrated in Fig. 4 ready for the beginning of a compression stroke.
  • a relatively high level of compression accumulator pressure P c is desirable.
  • the reset actuator 82 provides the necessary force level for effecting the desired resetting of the pistons.
  • the reset actuator 82 is a form of hydraulic pressure intensifier including an actuator piston 90 movable in a stepped cylinder 91 and having a relatively large surface area 92 exposed in a fluid chamber 93 and a relatively small surface area 94 exposed in a fluid chamber 95.
  • a spring 96 ordinarily biases the actuator piston 90 to a righthand position (not shown) in the stepped cylinder 91 when out of use.
  • a fluid flow path 97 through the actuator piston 90 and a check valve 98 therein provide unidirectional fluid flow coupling of the fluid chambers 93,95.
  • the reset actuator 82 is sized so that the total displaceable volume of the second pump chamber 61 is somewhat less than the total displaceable volume of the fluid chamber 95.
  • the large surface area 92 of the actuator piston 90 is greater than that of the small surface area 94 by an amount which is adequate to overcome the load of the spring 96, frictional forces, and the force due to pressure in the second pump chamber 61 to reset the pump piston 42 to the position illustrated in Fig. 4, i.e. against the compression force, i.e. the product of P c times the area of the compression piston 46.
  • the reset valve 78 When resetting occurs, the reset valve 78 is positioned as shown in Fig. 4. The first pump chamber 60 is then vented to low pressure P, and the reset actuator 82 is supplied with high pressure P o . The actuator piston 90 is driven toward the lefthand position shown in Fig. 4 forcing fluid from the fluid chamber 95 into the second pump chamber 61 driving the pump piston 42 and engine piston 5 to the righthand position shown in Fig. 4.
  • the pump piston 42 has fully reset, as may be sensed by a position sensor described in detail below with reference to Fig. 13, for example, the spool of the reset valve 78 is moved in its cylinder to the right blocking communication with the first pump chamber 60 and venting the reset actuator fluid chamber 93 to low pressure P,.
  • the fluid flow path 97 and check valve 98 then permit the spring 96 to move the actuator piston 90 to its maximum righthand position while the fluid chamber 95 is refilled with fluid and the check valve 83 isolates the pump from the chamber 95.
  • the FPEP 1 In the reset position illustrated in Fig. 4 the FPEP 1 is ready for starting cycle initiation. Moreover, the illustrated position and setting of the several control valves 67-69 is the "hold" condition between cycles when the FPEP 1 is operated in an intermittent manner.
  • control valve 67 is opened quickly thereby venting the high pressure in the second pump chamber 61 to low inlet pressure level, whereupon compression energy from the compression accumulator 49 effects a compression stroke commencing cyclical operation in the primary mode.
  • Such increase in compression energy may be accomplished initially by raising pressure P c in the hydraulic accumulator 49, and particularly of the fluid (preferably a compressible gas) in the bladder 50, to a predetermined level followed by the above-described resetting sequence and operation initiation with an initial compression stroke.
  • the control valve 69 may be opened during the initial compression stroke and is left open until it is desired to effect operation specifically in the primary or secondary modes described above. After normal operation is under way and the FPEP 1 is satisfactory warm, the compression accumulator pressure P c may be reduced somewhat to lower unnecessary compression energy losses.
  • the above noted resetting would be effected initially. Thereafter, to initiate a compression stroke, while maintaining the control valve 67 closed, the control valve 68 first is opened, and promptly thereafter the control valve 69 is opened.
  • the FPEP 1 accordingly would be configured for operation in the secondary mode of operation for succeeding cycles, as is described above with reference to Figs. 3A and 3B.
  • the FPEP 1 can be made to vary its cycling frequency from a maximum down to rates as low as a few cycles per minute. Such cycling frequency control is accomplished by interrupting the normal cycling motion at the end of a power stroke with a pause period. Each cycle in itself is a full velocity cycle in both compression and power stroke directions. Interruption occurs at the end of a power stroke to create a pause period until the interruption is terminated, and the interruption process is brought about by valving operation, as will now be described.
  • the inlet check valve 63 is active or passes fluid on the power stroke to provide a supply into the second pump chamber 61, and, therefore, the control valve 67 does not have to close particularly fast during the power stroke, although its closure should be completed by the end of the power stroke.
  • both control valves 68 and 69 must be sequentially actuated to open the same. More particularly, on a power stroke the control valves 68 and 69 are closed and fluid fills the second pump chamber 61 via the inlet check valve 63 and exits the first pump chamber 60 at high pressure through the outlet check valve 65. Compression stroke motion will be prevented by the pressure lock formed in the second pump chamber 61. To initiate the next cycle of operation, control valve 68 is first opened followed by opening of control valve 69 to initiate a compression stroke.
  • the acceleration booster 100 includes two substantially identical portions 100A, 100B shown in operative condition in Fig. 5A.
  • the acceleration booster 100B for example, includes a boost piston 101 slidable in a cylinder 102.
  • output pressure at port 105 acting on the exposed surface area of boost piston 101 in chamber 106 moves the boost piston to the lefthand position shown in Fig. 5A.
  • a pad 110 on the back side of the engine piston 5 engages the rod 111 of the boost piston 101 driving the same to the right causing an outflow of high pressure fluid from the chamber 106 via the port 105.
  • the engine piston 5 Since the surface area of the boost piston 101 exposed in chamber 106 is larger than the surface area of the rod 111 exposed in the back side chamber 31 of the engine portion 2, the engine piston 5 will decelerate to zero more rapidly than would be expected to occur without the acceleration booster 100 activated.
  • the area 107 times the pressure in the chamber 106, i.e. output pressure P o will be the force tending to expedite such deceleration.
  • the same pressure force provides quick acceleration during the beginning of the subsequent compression stroke. The net result is that both power and compression stroke time periods are shortened and the resultant cycle and delivery rates are increased.
  • the valve 103 ports the chamber 104 to output pressure P o urging the boost piston 101 to the right in its cylinder 102 until it engages the energy absorber piston 112.
  • the area of the boost pistion 101 exposed in chamber 104 is slightly larger than the area 107 exposed in chamber 106 so that the boost piston has adequate retraction force to the position shown, for example, in Fig. 5B but does not compress the heavy absorber piston spring 113.
  • FIG. 5B there is illustrated one of the portions 100B of the acceleration booster 100 of Fig. 5A along with the energy absorber 114, including the absorber piston 112 and spring 113.
  • the absorber 114 also includes a fluid-tight cylinder 115 in which the spring 113 is contained and the absorber piston 112 may slide.
  • a fluid passage 116 and check valve 117 are contained in the absorber piston 112, and a fluid path 118 between the chamber 119 in the cylinder 115 and the output port 105 conducts fluid therebetween as permitted by the absorber piston 112.
  • each energy absorber 114 also includes a high force liquid spring to decelerate the piston masses and absorb the excessive energy.
  • the absorber pistons 112 are closely fit to their respective cylinders 115.
  • the check valve 117 and fluid passage 116 ensure that all air is removed from the liquid spring chamber 119.
  • the fluid path 118 includes an orifice 120 closed by the absorber piston 112 promptly after retraction motion of the latter commences.
  • a plurality of FPEP's may be combined to increase power output and to add flexibility of operation. If desired, only a single FPEP of a combined group may be operated at a time, for example when hydraulic demands are low, or all of the FPEP's may be operated.
  • a grouped, here paired, FPEP system 130 includes a FPEP la' and a FPEP 1b', each of which is substantially the same in form and operation as the FPEP 1 described above.
  • primed reference numerals designate parts having the same or similar form and function as those designated by the same unprimed reference numerals in Fig. 1.
  • FPEP's When pairing FPEP's in accordance with the invention, it is desirable to combine some specific pump elements and functions to improve pumping efficiency and to reduce pressure pulsations normally associated with piston pumps.
  • the high and low pressure accumulators 53', 64' are shared to reduce flow losses and space requirements; as a result there is a net gain in output efficiency over a single FPEP.
  • the FPEP's 1a', 1b' are positioned about respective centerlines 131a, 131b, which are parallel and for the sake of clarity, the line 131b appears at the top and bottom of Fig. 6, with the pump piston and chambers of the hydraulic pump portion 3b' thereof being divided as shown.
  • the FPEP's la', Ib' have interacting elements, valving and porting between the two pumps in a side-by- side installation, and such FPEP's are made preferably to cycle alternately by the electronic control system described below with reference to Fig. 13.
  • control valves 67a', 67b' interconnecting the second pump chambers 61a', 61b' allowing fluid to pass freely between such second pump chambers with the negligible pressure losses.
  • control valves 67a', 67b' function to isolate the FPEP's 1 a', 1 b' when acting independently as well as to provide a conversion between primary and secondary modes of operation.
  • the inlet and outlet flow rates are essentially continuous if power stroke time approximates the compression stroke time. Flow will, of course, stop momentarily at the ends of the strokes at which time the accumulators 53', 64' supply the flow demand.
  • the system 130 has the control valves set for operation in the primary mode. The operation insofar as intermittent cycling with a pause period, as was described above, the starting, and the general operation of the system 130 using both FPEP's 1a', 1b' operating out of phase with each other or using only one of them at a time will be substantially the same as is described above, and, if desired, the acceleration boost and energy absorbing features described above also may be included in the system 130.
  • system 130 may be operated in the manner described above in the secondary mode, for example, by opening the ocntrol valves 68a', 68b', 69a', 69b', and closing the control valves 67a', 67b'.
  • control valve 67b' would be closed while the control valve 67a' remains open to permit operation, say in the primary mode, or the FPEP 1 b' is operated in the secondary mode, as was described above.
  • Cylinder air compression requirements are well known by those involved in the design of diesel engines.
  • the primary purpose of compressing the cylinder air charge is to increase its pressure and associated temperature to an adequate level for ignition of the diesel fuel when it is injected. This requirement varies depending on conditions such as the initial temperature of the cylinder and air and pressure level of the inlet air charge.
  • the final condition of the air charge in the cylinder when compressed is directly related to the initial cylinder volume at the point of inlet port closing divided by the final cylinder volume. This is termed the compression ratio of the engine and is generally fixed in the rotary engine design or variable in some specific applications by mechanical means or by a limited hydraulic control means in the piston proper.
  • the mechanizations to date have been limited in flexibility and have generally been designed for specific purposes such as to reflect peak cylinder pressures and avoid structural failures or for experimental purposes.
  • the opposed piston free piston engine pump 1 disclosed has unique capabilities and flexibilities in this area. They are the provision of total control over the amount of energy put into compression to establish compression ratio and related pressure and temperature conditions within the cylinder, thereby optimizing overall efficiency of the system, minimizing compression losses, and controlling operating pressure profile within the cylinder, and the provision of control over the rate at which the compression energy is applied which establishes the velocity and acceleration profiles of the piston on the compression stroke and results in cycle rate variability.
  • Typical diesel engine cylinder characteristics and their relationship to compression ratio are shown in Figs. 7-9. These are shown as background information to establish the compression energy requirements and demonstrate the flexibility of the FPEP disclosed.
  • Fig. 7 shows the relationship between compression ratio and cylinder gas temperature for cold start-up of a typical diesel cylinder.
  • the opposed piston FPEP 1 is designed to achieve an equivalent cold start compression ratio in the range of 20 to 30 or higher.
  • Fig. 9 shows a plot of characteristic peak cylinder combustion pressure that can be expected vs. compression ratio for various brake mean effective pressures. Typical FPEP operating lines are shown. The characteristic BMEP lines show dramatically that compression ratio has a great impact on cylinder peak pressure level.
  • the opposed piston FPEP 1 of the present invention compression energy implementation controls the peak pressure levels within the limitations of the design while maintaining high thermal efficiencies. Extremely high peak cylinder pressures can readily be supported by the FPEP 1.
  • the combustion chamber is structurally suited for high pressure containment. Piston forces are transferred directly into hydraulic forces and acceleration of the piston elements. No crank arm or other linkage exists to resist the high acceleration forces on the piston.
  • Fig. 10 shows a typical FPEP plot of compression energy required vs. compression ratio for various cylinder air charge pressures.
  • the compression energy is stored as compressed gas in the compression accumulator 49 (Fig. 1).
  • the amount of energy available for compression is approximately defined by the following relationship:
  • FIG. 10 An operating line has been added to Fig. 10 showing a typical optimized control condition for best overall efficiency. This line is established by actual test results obtained for a specific engine application. Also shown in Fig. 10 is the cold start capability region of the design.
  • the energy that is available for compression and is stored in the compression accumulator is shown in Fig. 11.
  • the characteristic shown is based on an accumulator gas charge volume of 492 cm 3 (30 in3) and pre-charge pressure of 70 Bars (1000 psi).
  • the working displaced volume has been selected as 80 cm 3 (5 in3).
  • a pressure control valve 70 as shown in Fig. 1 establishes the nominal pressure level of the compression accumulator. This is varied during operation so that the predetermined performance requirements are achieved.
  • the pressure control valve 70 receives its information from the electronic microprocessor control centre as described below.
  • the compression energy is varied by as much as 3 to 1 or more, as shown in Fig. 11 for example. This adequately covers the operating requirements indicated in Fig. 10 which varies by approximately 2 to 1.
  • some of the energy for compression is supplied by the output pressure accumulator 53 as explained earlier.
  • the amount of energy required of the compression accumulator 49 is, therefore, decreased toward the low energy range shown in Fig. 11.
  • Fig. 12 shows how the pressure force applied to the piston mass (42, 46, 5) varies with stroke and energy level. As the net energy level is raised by increasing the working pressure of the compression accumulator 49, the initial force at the start of the compression stroke increases significantly with respect to the final force level. The advantageous result of this profile change is that the initial acceleration of the piston mass increases substantially. This provides an effective acceleration control means for "speeding up" the compression stroke time and resultant cycle rate.
  • Fig. 13 the basic elements of the primary control circuit 150 are illustrated in association with the FPEP 1. Only primary inputs to the microprocessor electronic control center 154 are indicated along with the outputs that control FPEP operation. Other inputs to the electronic control 154 of secondary importance also may exist; these would be expected usually to have only low priority influence on the output signals.
  • the secondary inputs include such information as intake and exhaust manifold temperatures, exhaust pressure, other oil pressures and temperatures, and failure detection sensors.
  • the primary control loop accomplishes the following functions.
  • the primary control loop provides means of regulating the cycle-by-cycle operation of the FPEP 1 to achieve consistent operation and optimized performance potential for all hydraulic pressure and flow demands within its design range.
  • the delivery of the FPEP may be controlled in an efficient manner including cycle rate variability with turn-down ratio of 100 to 1 or greater.
  • the compression energy and resultant compression ratio required by the engine combustion process may be controlled to achieve the most efficient operating potential over all power output levels including inlet air supercharge pressure levels of three atmospheres and greater.
  • the control loop provides means of quickly positioning the engine- pump elements for start-up as well as a method of controlling the dual pumping feature of the pump to provide a smooth transition between primary high flow low pressure mode of operation and secondary low flow high pressure mode of operation.
  • the compression energy may be substantially increased at the first part of the compression stroke to accelerate the cycle rate and increase the pumping rate capability of the FPEP 1.
  • the FPEP module cycle rate may be synchronized with that of adjacent FPEP module cycle rates for the purpose of providing continuous hydraulic input and output flow.
  • the opposed piston FPEP 1 has the flexibility of varying the bottom dead position (start of compression stroke) as required to optimize the energy output process.
  • Fig. 14 indicates that a control range exists around the nominal bottom dead position of the stroke for various power levels and cycle rates when optimizing overall performance.
  • the interrelationship of factors that shape the combustion gas diesel cycle such as intake and exhaust port areas, inlet air pressure level and output power level can be optimized by varying the bottom dead position as determined by actual test data and programmed into the logic of the electronic microprocessor control 154.
  • piston position, inlet air charge pressure, and output level are the primary engine sensor inputs 151-153 to the electronic control 154. Based on the predetermined algorithm for these system parameters, the injector fuel delivery control 155 operation is varied as required to establish the required operational stroke length.
  • hydraulic power output from the FPEP 1 is regulated by input command control 157 and output pressure level rate of decay information sensed by pressure sensor 153.
  • the electronic control 154 determines the cycle rate, fuel delivery setting, and air charge pressure required to meet the need based on the rate of pressure level decay by controlling the fuel delivery control 155 and fuel injector 10, the desired supercharging, and the pump control valves 160 (such as the valves 67-69, 78, and 103 of Figs. 1, 4 and 5A).
  • Compression energy level may be controlled based on pre-programmed information stored in the electronic control 154 in order to adjust the pressure level of the compression accumulator 49 for optimized compression.
  • the sensor 161 senses compression accumulator pressure and this is correlated with the information received by the electronic control 154 from the input command 157, inlet air pressure sensor 152 and output pressure sensor 153 to operate the pressure control valve 162 which either raises or lowers the pressure level as required for the particular mode of operation.
  • the electronic control 154 logic determines whether the FPEP 1 should be started based on information including operator input 157, output 153 and compression 161 accumulator pressures, inlet air pressure 152 and piston positions 151. If conditions are satisfied, the FPEP pistons 5, 6 will be reset, the reset valve 78 will be closed, and the start cycle will be initiated.
  • Mode selection can be made by either external operator input 157 or by automatic electronic control via sensors 152, 153 161 depending on hydraulic pressure/flow load requirements.
  • output 53 and compression 49 accumulator pressures are sensed and inlet air charge pressure to the combustion chamber 7 must be read; necessary adjustments are made prior to switching the control valves 67-69 by the electronic control 154, for example, which also preferably adjusts the compression accumulator 49 pressure and appropriately controls the fuel injector 10 to accomplish change and mode of operation.
  • Two or more FPEP modules 1 within an installation may be synchronized by slight piston stroke length changes in one module as compared with another as reference.
  • the piston stroke length 151 is decreased or increased as required to change the cycle rate of the module in tune with the reference module 170.

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  • Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Reciprocating Pumps (AREA)
  • Control Of Positive-Displacement Pumps (AREA)

Claims (10)

1. Einbaupumpsystem mit Gegenkolben-Freikolbenmotor, mit einem Freikolbenmotor (2), welcher einen Motorzylinder (4) mit linearer Achse einschließt, einem Paar von Motorkolben (5,6), welche in dem Motorzylinder entlang einer solchen Achse während eines Kompressionshubes gegeneinander und während eines Arbeitshubers auseinander bewegbar sind; einem getrennten Pumpmittel (3), welches mit jedem Motorkolben zum Pumpen von Fluid verknüpft ist, wobei jedes Pumpmittel einen Pumpenzylinder (43) einschließt, welcher axial mit dem Motorzylinder ausgerichtet ist, einem Pumpkolben (42) welcher mit dem entsprechenden Motorkolben für die Bewegung mit diesem gekoppelt ist und erste und zweite Pumpenkammern (60, 61) innerhalb des Pumpenzylinders bildet, Ventilmitteln (67, 68, 69), welche für die Steuerung des Flusses von Fluid in die Pumpkammern (60, 61) hinein und aus diesen heraus vorgesehen sind, wobei die Ventilmittel ein erstes Absperrventilmittel (62) zum Durchlassen eines Einlaßflusses von Fluid unter relativ geringem Druck in die erste Pumpkammer (60) aufweisen, sowie ein zweites Absperrventilmittel (65) zum Durchlassen des Auslaßflusses von unter relativ hohem Druck stehendem Fluid aus der ersten Pumpkammer, dadurch gekennzeichnet, daß ein erstes, wahlweise betreibbares Ventilmittel (69) parallel zu dem zweiten Absperrventilmittel (65) angeordnet ist, um letzteres wahlweise zu umgehen, dritte und vierte Absperrventilmittel (63, 66) zum Durchlassen des Einlaßflusses von Fluid unter relativ niedrigem Druck in die zweite Pumpe bzw. von Auslaßfluß von Fluid unter relativ hohem Druck aus der zweiten Pumpkammer, ein zweites, wahlweise betreibbares Ventilmittel (67), welches parallel mit dem dritten Absperrventilmittel (63) angeordnet ist zur Umgehung des letzteren, und ein drittes, wahlweise betreibbares Ventilmittel (68), welches in Reihe mit dem vierten Absperrventilmittel (66) angeordnet ist, um wahlweise den Fluidfluß durch das letztere zu steuern.
2. System nach Anspruch 1, weiterhin dadurch gekennzeichnet, daß das Ventilmittel flußmäßig mit dem Absperr- bzw. Rückschlagventilmittel und den Pumpenkammern verbunden ist, um den wahlweisen Betrieb der Pumpmittel in einer Hochdruckbetriebsart und in einer Hochflußbetriebsart zu ermöglichen, und Einlaß-Fluidfluß-Mittel (Prn) vorgesehen sind, um an die Pumpmittel eine Fluidzufuhr mit relativ niedrigem Druck anzuschließen, wobei die Einlaß-Fluidfluß-Mittel so verbunden sind, daß sie eine derartige Fluidzufuhr zu dem ersten Absperr- bzw. Ruckschlagventilmittel (62) und zu dem parallel angeschlossenen dritten Absperr- bzw. Rückschlagventilmittel (63) und dem zweiten wahlweise betreibbaren Ventilmittel (67) vorsehen, und ein Auslaß-Fluidfluß-Mittel (P""t) um das zweite Absperrventilmittel (63) und das erste wahlweise betreibbare Ventilmittel (6) und das in Reihe geschaltete vierte Absperrventilmittel (66) und das dritte wahlweise betreibbare Ventilmittel (68), (welche parallel angeschlossen sind), miteinander und mit einem unter relativ hohem Druck stehenden Fluidauslaß zu verbinden.
3. Das Freikolbenmotor-Pumpsystem nach Anspruch 1, weiterhin dadurch gekennzeichnet, daß Sensormittel (151-153) zumindest einen Parameter der Motor- und Pumpmittel überwachen und daß ein elektronisches Steuermittel (154), welches auf die Sensormittel anspricht, wenigstens eine Wirkungsweise des Motors und des Pumpmittels steuert.
4. System nach Anspruch 3, weiterhin dadurch gekennzeichnet, daß wahlweise betreibbare Zusatzbeschleunigungsmittel (100) vorgesehen sind, für die Erhöhung der anfänglichen Beschleunigung der Motorkolben bei einem Kompressionshub, mit einem Zusatzkolbenmittel (101), welches mit den Motorkolben in Eingriff bringbar ist, um während eines Arbeitshubes in eine Richtung bewegt zu werden und in die entgegengesetzte Richtung während des Beginns eines Kompressionshubes, um den Motorkolben eine Arbeit zuzuführen, welche dazu führt, daß die letzteren bewegt werden, um eine Kompression zu bewirken, und dadurch der Kompressionshub beschleunigt wird, und daß das Fluidsteuermittel (103) wahlweise betreibbar ist, um die Bewegung des Zusatzkolbenmittels in eine aktivierte Position zu bewirken, um eine solche erhöhte Beschleunigung hervorzurufen und in eine entaktivierte Position, in welcher das Zusatzkolbenmittel normalerweise nicht mit den Motorkolben in Eingriff steht.
5. System nach Anspruch 4, weiterhin dadurch gekennzeichnet, daß Energieabsorbermittel (14) vorgesehen sind zum Absorbieren überschüssiger Energie, welche durch einen Überweg der Motorkolben erzeugt wird, mit einem Energieabsorberkolben (112), welcher mit dem Zusatzkolbenmittel (101) in Kontakt bringbar ist, wenn das letztere entaktiviert oder in seinem aktivierten Betriebszustand ist und von den Motorkolben zumindest um eine vorbestimmte Strecke bewegt wurde.
6. Das Freikolbenmotor-Pumpsystem nach Anspruch 1, weiterhin dadurch gekennzeichnet, daß Synchronisationsmittel (16) vorgesehen sind zur Aufrechterhaltung einer synchronen Bewegung der Motorkolben normalerweise ohne wesentliche Kraft zwischen ihnen zu übertragen, um dadurch im wesentlichen relativ zum Schwerpunkt des Motors im wesentlichen ausbalanciert zu bleiben.
7. Das Freikolbenmotor-Pumpsystem nach Anspruch 6, weiterhin dadurch gekennzeichnet, daß das Synchronisierungsmittel einen Zahnstangen-und Ritzelaufbau (17-24) aufweist.
8. Das Freikolbenmotor-Pumpsystem nach Anspruch 1, weiterhin dadurch gekennzeichnet, daß eine weitere Freikolbenmotorpumpe (1 b'), welche Einlaß- und Auslaß- Fluidfließleitungen hat, parallel mit der ersten Freikolbenmotorpumpe (1a') verbunden ist, wobei die weitere Freikolbenmotorpumpe umfaßt: einen Motorzylinder (4) mit einer linearen Achse, zwei Motorkolben (5, 6), welche in dem Motorzylinder entlang dieser Achse während eines Kompressionshubes gegeneinander und während eines Arbeitshubes voneinander wegbewegbar sind, und ein separates Pumpmittel (3), welches mit jedem Motorkolben verknüpft ist, um Fluid zu pumpen; und daß ein Steuermittel (150) zum Steuern des Betriebes der Freikolbenmotorpumpen vorgesehen ist.
9. Freikolbenmotor-Pumpsystem nach Anspruch 1, weiterhin dadurch gekennzeichnet, daß ein Druckspeicher (49) vorgesehen ist, mit einem relativ festen Gehäuse (51), Fluid-Anschlußmitteln (48), um ein erstes Fluid in das Gehäuse und aus dem Gehäuse heraus anzuschließen, einem verformbaren Teil (50) in dem Gehaüse, welches einen im wesentlichen fluiddichten Behälter darin bildet, und einem Medium in dem Behälter, welches zusammenpreßbar ist, um Energie zu speichern in Abhängigkeit von dem auf den Behälter durch das erste Fluid während eines Arbeitshubes aufgebrachten Druck, und welches ausdehnbar ist, um den Behälter auszudehnen und dadurch Energie an das erste Fluid abzugeben, um einen Kompressionshub zu bewirken.
10. Das Freikolbenmotor-Pumpsystem nach Anspruch 1, weiterhin dadurch gekennzeichnet, daß Auslaßöffnungsmittel (12) vorgesehen sind in dem Motorzylinder (4) in der Nähe eines Endes desselben, bezogen auf die normalerweise erwartete maximale Auslenkung eines Motorkolbens während eines Arbeitshubes, um durch eine solche Auslenkung geöffnet zu werden, um Ausstoßprodukte der Verbrennung aus dem Motorzylinder herauszulassen, und Lufteinlaßöffnungsmittel (11) in der Nähe des entgegengesetzten Endes des Motorzylinders, bezogen auf die normalerweise erwartete maximale Auslenkung des anderen Motorkolbens während eines Arbeitshubes, um durch den letzteren Motorkolben geöffnet zu werden, um Luft für die Verbrennung in dem Motorzylinder zuzuführen, wodurch die Öffnungsmittel für ein Spülen in einer Richtung in dem Motorzylinder sorgen.
EP19820300655 1982-02-10 1982-02-10 Einheit aus Freikolbenmotor und Pumpe Expired EP0085800B1 (de)

Priority Applications (2)

Application Number Priority Date Filing Date Title
EP19820300655 EP0085800B1 (de) 1982-02-10 1982-02-10 Einheit aus Freikolbenmotor und Pumpe
DE8282300655T DE3276338D1 (en) 1982-02-10 1982-02-10 Opposed piston type free piston engine pump unit

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
EP19820300655 EP0085800B1 (de) 1982-02-10 1982-02-10 Einheit aus Freikolbenmotor und Pumpe

Publications (3)

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EP0085800A2 EP0085800A2 (de) 1983-08-17
EP0085800A3 EP0085800A3 (en) 1984-07-25
EP0085800B1 true EP0085800B1 (de) 1987-05-13

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Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2020044176A1 (en) * 2018-08-31 2020-03-05 Narasimha Murthy P L Ahobala Internal combustion engine with split cylinder and free piston and power generation using the same

Families Citing this family (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE4024591A1 (de) * 1990-08-02 1992-02-06 Gerhard Brandl Freikolbenmotor
DE10120196A1 (de) * 2000-05-19 2001-11-22 Mannesmann Rexroth Ag Freikolbenmotor
DE10033443A1 (de) * 2000-07-10 2002-02-07 Mannesmann Rexroth Ag Freikolbenmotor

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Publication number Priority date Publication date Assignee Title
FR1350876A (fr) * 1962-12-21 1964-01-31 Anciens Etablissements Panhard Perfectionnements apportés aux dispositifs de synchronisation des moteurs thermiques à deux cylindres et à deux pistons opposés dans chaque cylindre
US3432088A (en) * 1967-05-24 1969-03-11 Sulzer Ag Free piston-type internal combustion pumping engine
NL160632C (nl) * 1968-10-08 1979-11-15 Ir Theodorus Gerhardus Potma Vrije-zuigerpompinstallatie.
GB1332799A (en) * 1970-10-12 1973-10-03 Riekkinen As Hydraulic power unit including a hydraulic pump operated by a free piston internal combustion engine
US4307999A (en) * 1979-06-25 1981-12-29 Pneumo Corporation Free piston engine pump including variable energy rate and acceleration-deceleration controls
US4382748A (en) * 1980-11-03 1983-05-10 Pneumo Corporation Opposed piston type free piston engine pump unit

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2020044176A1 (en) * 2018-08-31 2020-03-05 Narasimha Murthy P L Ahobala Internal combustion engine with split cylinder and free piston and power generation using the same

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Publication number Publication date
EP0085800A2 (de) 1983-08-17
DE3276338D1 (en) 1987-06-19
EP0085800A3 (en) 1984-07-25

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