US3104801A - Infinite step controlling device for compressors - Google Patents

Description

Sept. 24, 1963 P. A. BANCEL 3,104,801

INFINITE STEP CONTROLLING DEVICE FOR COMPRESSORS Filed March 3, 1960 3 Sheets-Sheet 1 250 254 gm F/6'.6

INVENTOR PAUL A. BANGEL P. A. BANCEL 3,104,801

INFINITE STEP CONTROLLING DEVICE FOR COMPRESSORS Sept. 24, 1963 I5 Sheets-Sheet 2 Filed March 3, 1960 FIG. 7

TRAVEL OF VALVE CHANNELS IN INCHES INVENTOR ,3 PAUL A. BANOEL P 1963 P. A. BANCEL 3,104,801

INFINITE STEP CONTROLLING DEVICE FOR COMPRESSORS Filed March 3, 1960 3 Sheets-Sheet 3 I82 I V j 22 |9962 |s|\ 202 I98 we |90 88 I68 200 I80 6 e4 I70 I86O8 I64 F/G.8 INVENTOR PAUL A. BANGEL HIS ATTORNEY United States Patent 3,104,801 lNFE STEP CONTROLLING DEVICE FOR CGNEPRESSORS Paul A. Rance Montclair, NJL, assignor to lngersoll- Rand Company, New York, N.Y., a corporation of New Jersey Filed Mar. 3, 1969, Ser. No. 12,565 3 Claims. (Cl. 23025) The present invention relates to compressors and pertains more particularly to means for controlling the output of constant speed compressor to any range within its capacity.

The invention is particularly utile when applied to a compressor driven by a constant speed drive, such as a synchronous electric motor which is not readily adapted for varying the capacity of the compressor.

Heretofore, a preferred method for regulating the out put of a compressor has been by means of clearance pockets at the end of the cylinder. Each pocket is generally provided with a clearance valve, which opens into the cylinder bore. Thus when the clearance valve is opened, the volume of the pocket is added to the normal cylinder clearance volume. The effect of this is to decrease the amount of fluid taken into that end of the cylinder. More than one pocket is generally available :for conjunction with the cylinder bore. In this manner by regulating the number of pockets opened, the capacity of the compressor may be regulated from zero to full capacity. But it will be apparent to those skilled in the art, that such regulation can be attained only by fixed and definite increments, as each pocket is opened. Precise regulation to close capacity cannot be obtained.

Another method of securing capacity control to some extent is by the means of so-called free air unloaders. A free air unloader is a device which holds the inlet valve open to unload one end of a compressor cylinder end and thereby reduce its capacity to zero. But with this device too, only a limited amount of control is provided, because primarily the variation in capacity is determined by the number of cylinder ends available to unload.

In still another system the speed of the compressor is varied in response .to capacity demands, and since the volume of fluid output varies directly with the speed of the compressor, the actual output follows the demand closely. But this system is only applicable wherever the speed of the prime mover may be varied, and not to such compressors as driven by a synchronous electric motor, or any constant speed prime mover. Furthermore in the case where a compressor is being driven by a steam engine or a gas engine, it is not possible to drive the compressor to zero output because the revolutions per minute of the driving engine cannot be finitely controlled below a certain minimum value, beyond which stalling of the engine will occur.

Still other sysems consists of a combination of the foregoing systems and such arrangements are generally more expensive and require greater attention and maintenance care.

The present invention contemplates a novel valve control means for use in a control system based upon infinite step control whereby clearance pockets and their appended valves are eliminated while insuring precise and selective capacity control over the entire range of the compressor.

By the term infinite step control is meant simply that the point during the compression stroke of the piston at which the inlet valve is permitted to close may be precisely selected from any of an infinite number of points along the travel of the piston, so that compression of the fluid in each cycle will not begin until the piston reaches that point. Thus any undesired quantity of fluid will be expelled through the open inlet valve until the piston approaches the selected point.

This invention contemplates providing means for closing the inlet valve in a gentle yet positive manner without excessive strain and impact thereon. When the inlet valve on the compressor, in a system of this nature is held open, and the compressor piston begins the compression stroke, velocity through the valve builds up very rapidly against the valve to a relatively high magnitude. Thus the agency tending to hold the valve open is suddenly overcome and the valve is slammed shut with great force and violence. The fatigue stresses thus imposed upon the valve cause frequent and costly breakage with resultant operational failure of the entire system. The problem of keeping the inlet valve open against the powerful, velocity forces built up by the oncoming piston and then closing the valve at a slow and sluggish velocity to prevent excessive wear thereon is therefore of prime importance and one upon which the trouble-free operation of the system depends.

The present invention provides a solution to this heretotore unsolved problem by means of a hydraulic cushioning pocket against the end of the above mentioned ram to decelerate the inlet valve, and the invention further provides that said cushioning pocket shall be brought into operation against the ram by the very withdrawing motion of the ram itself as it retreats away from the inlet valve, and that the timing and magnitude of the cushioning force may be regulatively operated to insure the best performance.

A principal object of this invention is to provide means for closing the inlet valve of a compressor in a gentle yet positive manner.

A further object of the invention is to provide means 'for holding the inlet valve of a compressor open against the compression stroke of the compressor for a predetermined period, and then close the valve in a gentle and gradual manner to prevent breakage thereof.

The foregoing and further objects and advantages will appear more fully hereinafter from a consideration of the detailed description which follows, taken together with the accompanying drawings wherein one embodiment of the invention is illustrated by way of example. It is to be expressly understood, however, that the drawings are for the purpose of illustration only, and are not to be construed as defining the limits of the invention.

In the drawings wherein like reference characters indicate like parts;

FIGURE 1 is a diagrammatic view of the system showing the inter-relation of the various components associated therewith; 1

FIG. 2 is a sectional view of the pressure regulator (for the system;

FIG. 3 is a sectional view of the inlet valve of the compressor with the operating mechanism associated therewith indicating the cushioning means for said mechanism;

FIG. 4 is a section taken along the line 44 of FIG. 3

FIG. 5 is a section taken along the line 5--5 of FIG. 3;

FIG. 6 is a sectional view of the adjustable pressure relief valve for the system;

FIG. 7 is an enlarged view of the portion of FIG. 3 designated by the letter A;

FIG. 8 is a sectional view of the hydraulic pump for the system indicating the adjustable output means and pressure bypassing means associated therewith;

FIG. 9 is a section taken along the line 9--9 of FIG.

FIG. 10 is a section taken along the line 10-10 of FIG. 8;

FIG. 11 is a section taken along the line 1111 of FIG. 8;

FIG. 12 is a section taken along the line 12-12 of FIG. 8 indicating the output adjusting mechanism for the pump; and

FIG. 13 is a graph showing the relation between the closing velocity and travel of the valve operating mechanism shown in FIG. 3.

Referring now to the drawings and more particularly to FIG. 1, there is included in the infinite step control system, a reciprocating compressor gene-rally designated by the numeral it), driven by a rime mover 12. The compressor 16 may be any of a number of standard types and comprises a cylinder 14 with a piston 16 reciprocating therein, and an inlet valve 18 such as shown in detail in FIG. 3.

The prime mover 12 is actuated by a gas engine, generally designated herein by the numeral 20. It will be appreciated that the compressor may also be driven in any approved manner such as by an electric synchronous motor or the like. A piston rod 22 is connected to the prime mover 12 in a standard manner (not shown here) to supply motive power to the piston 16.

Connected also to the engine 20, and deriving operating power therefrom is avariable output pump 24- (FIGS.

1 and 8). The pump 24 is inter-connected with the engine 2% in a conventional manner, by means of a bevel gear arrangement 26 between a driving shaft 28 on the engine, and a driven shaft 39. A cam 32 on the shaft 3% translates the rotational motion of the driven shaft 3i) into a linear reciprocating motion to operate the pump 24. The more detailed operation thereof is hereinafter explained.

The pump 2 is'connected on the inlet end thereof to a hydraulic oil reservoir 34, and the outlet end is connected firs-t to a pressure relief valve, generally designated by the numeral 36 (FIG. 6) and thence to a hydraulic ram mechanism 38 (FIG. 3) for operating the cornpressor inlet valve 18. The more detailed operation of the mechanism 38 will be hereinafter explained.

The pump 24 is also in mechanical connection with a diaphragm motor 49 by means of a. rack and gear as se-mbly l2 (FIGS. 1 and 8), in such manner hereinafter explained, that the actual hydraulic oil output of the pump may be regulated by the movement thereof.

The diaphragm motor 40 is connected and responsive to the pressure in the compressor receiver 4-4. However, an intermediately positioned pressure regulator generally designated by the numeral 46 (FIGS. 1 and 2) governs the actual pressure transmitted to the diaphragm motor 40.

Hence it will be appreciated that generally expressed, the manner in which the herein disclosed infinite step control system operates is as follows:

The engine 20 drives the compressor 10 and the pump 24 at a constant speed. The compressor 1% therefore delivers a quantity of compressed fluid to the receiver 44. As the demand on the receiver 44 varies due to consumption of the fluid by tools or instruments, the variation in demand is reflected in a pressure change which is signaled to the pressure regulator 46. The pressure regulator now transmits the new pressure to the diaphragm motor 48. The diaphragm motor 40 reacts to the pressure so transmitted in direct proportion to its magnitude on the rackand gear assembly 42, which in turn fixes the output of the pump 24. If the pressure in the receiver has been lowered, it follows that greater fluid output from the compressor is required to restore its value. Conversely, if the pressure has increased, less compressor output is desirable. The operation of the inlet valve 18 is controlled by the output of hydraulic oil from the pump 24. The inlet valve 18 as is understood, governs the output of the compressor 16 by regulating the quantity of fluid permitted into the cylinder 14 for compression purposes. Therefore, the pump 24 4 now responsive to the new setting operates the inlet valve'18. inlet valve will be held open even though the piston 16 is well into its compression stroke. portion of the fluid will be expelled from the compressor back through the inlet valve hence its output will decrease. Conversely if the output of the pump 24 has been decreased, then the inlet valve 18 will close sooner and a greater amount of fluid than otherwise, will be 7 compressed to augment the output of the compressor. it will now be appreciated, that in the foregoing manner it is possible to control the output of the compressor throughout the entire range of its capacity, from zero percent to one hundred percent by closing the inlet valve 18 as the piston 16 reaches any selected one of an infinite number of points along its compression stroke.

Pressure Regulator Operation Referring now to FIG. 2 for a more detailed consideration of the pressure regulator 46, a substantially bell shaped casing 48 contains within its confines a pressure chamber 59 open to the pressure in the receiver 44 by means of a conduit 51 and an inlet port 52. A second pressure chamber 54 is separated from the chamber 50 by an interior rib 56 having a concentric threaded boss 58 into which a needle valve assembly 69 is matingly threaded. A shoulder 62 thereon compresses a sealing ring 64 against the boss 58 to prevent any pressure seepage between the two chambers. A valve stem 66 within the assembly 60 extends through an opening 68 into the chamber 59 and is tapered on one end to completely close the opening 68 under the urging of a coilspring 7t compressed between an abutment 72 integral with the tapered end of the valve stem and a valve seat 74. The un-tapered end of the valve stem 66 extends through an opening 76 in the valve seat 74 which communicates with the second pressure chamber 54. 'Ihus the two pressure chambers 50 and 5d may be associated when the force of the coil spring 70 is overcome and the tapered end of the valve stem 66 unobstructs the opening 68. The upper portion of the casing 48 comprises simply means for opera-ting the valve assembly 60, in the form of a cylindrical chamber 78 which contains a powerful coil spring 88 anchored on one end against of the pressure chamber 50 and hermetically seal it from the cylindrical chamber 78. A concentrically positioned shaft 90 having a relatively wide base 92, extends through the plate 82 and the diaphragm 88 to bear against the valve stem 66 of the needle valve assembly 60. A looking nut 94 hence draws the base 92 tightly against the diaphragm 88 to lock it against the plate 82, thus pro-.

viding sealing means for the chamber 50 and securing the shaft 90 to the pressure sensitive diaphragm .88. Hence the spring continuously biases the tapercdvalve' stem d to free the opening 68. A fixed orifice 91'communicating with the ambientatmosphere provides a con stant set bleed for fluid pressure from the chamber 54 so that ressure therein cannot build up sufiiciently to equalize the pressure in the chamber 58, therefore insuring that only the needle valve 60 controls the differential in pressure between the two chambers.

Thus with an inter-relation and positioning of parts and components as herein described, it will now be appreciated that if relatively high fluid pressure from the receiver 44 compared to the setting of the spring 8.0 is

admitted into the chamber 50 through the inlet port 52 and exerted against the base 92 and the diaphragm 88,

If the pum output has been increased, the

In this manner a i the coil spring 80 will be further compressed. As the diaphragm rises in response to the fluid pressure, the valve stem 66 is biased towards closing the opening 68 by force of the spring 70. Hence the fluid pressure passing through the needle valve assembly 60 from the pressure chamber 50 into the pressure chamber 54 will be limited towards a lower value. Conversely when relatively lower pressure than that determined by the setting of the coil spring 80 is introduced into the pressure chamber 50, the coil spring will operate to force the diaphragm 88 and shaft 90 downward into the chamber 50, and the base 92 hearing on the end of the valve stem 66 will force the valve stem axially downward to open the needle valve assembly 68 thereby introducing more fluid pressure into the pressure chamber 54. It is to be noted that the value of the fluid pressure introduced into the pressure chamber 50 from the receiver 44 is an inverse indication of the output required from the compressor at any given time. It will further be appreciated that it is desirable to maintain the pressure in the receiver 44 at near constant pressure. An outlet port 95 is provided in the pressure chamber 54 to discharge the fluid pressure therein through a conduit 96 into the diaphragm motor 40.

Diaphragm Motor Operation The diaphragm motor 48 (FIG. 1) is any of a number of standard types commonly employed in the art for applications where it is required to translate a fluid pressure value into the finite linear motion of a prime mover with a certain force and amplitude directly proportional to said fluid pressure value. In the embodiment herein indicated, the diaphragm motor 40 comprises a casing 98 having an oval section 100 in one end and a cylindrical section 1132 in the opposite end. A pressure responsive membrane 184 extends yieldingly across the oval section 188 to define a separate pressure chamber 106 therein. The chamber 186 receives and contains therein, by means of the conduit 96 and an inlet port 110, the

fluid pressure from the pressure chamber 54 of the pressure regulator 46, and further is hermetically sealed from a chamber 108 by means of the membrane 184. Hence the fluid pressure discharged from the regulator 46 is brought to bear directly against the membrane 104.

The cylindrical section 102 contains therein a concentrically disposed coil spring 112 anchored on one end against the bottom of the cylindrical section 102 (FIG. 1) and on the other end against a spring retainer 114 which is integrally attached to a motor shaft 116. The motor shaft 116 extends concentrically within the diaphragm motor 48, and one end thereof is perpendicularly attached to the membrane 184 by means of a Washer 118 on both sides of the membrane and a retaining nut 128 threaded into the end of the motor shaft. The opposite end of the motor shaft 116 extends through the bottom of the cylindrical section 182 to engage and threadably attach to the rack and gear assembly 42 on the pump 24. The spring retainer 114 positioned in the opposite end of the cylindrical section 102 serves also to support the motor shaft 116 and to maintain the axial alignment thereof.

In the simple operation of this device it will now be apparent that when the fluid pressure is introduced into the chamber 186, the force exerted against the pressure responsive membrane 184 as a result thereof will motivate the motor shaft 116 against the force of the coil spring 112 to compress the same and lineally operate the rack and gear assembly 42. When the fluid pressure in the chamber 106 falls below the return force stored in the coil spring 112, the spring will resile and force the motor transmitted thereto from the pressure regulator 46 responding to the fluid pressure status of the receiver 44 to thereby indicate the output required of compressor 10.

Pump Operation Referring now to FIG. 8 for a more detailed consideration of the pump 24, a substantially cylindrical housing 122 concentrically encloses a plunger casing 124 having a longitudinal bore 126 adapted to contain a plunger 128. The plunger 128 not only is adapted to reciprocate in the bore 126 but is also adjustably rotatable therein as presently explained. A rotary gear 130 envelopes the plunger casing 124 (FIGS. 8 and 12), the gear teeth 132 of which mesh with teeth 134 on a linear rack 136. Hence linear motion of the rack 136 will result in rotary motion of the gear 130. The gear 130 is integral with a cylindrical section 138 which envelopes the plunger casing 124 and extends downwardly beyond the plunger casing in a concentric manner with the plunger 128. The aforementioned cylindrical section 138 has a pair of longitudinal slots 140, cut from the end thereof and diametrically positioned (FIGS. 8 and '11), adapted to mate with and slidably receive a key lever 1 42. The key lever 142 cucompasses the plunger 128 and is attached thereto as by a friction screw 144 (FIG. 11) so that as the plunger 'reciprocates in the bore 126, the key lever 142 rides attached thereto in the slots 148*. It is perceivable, that linear motion of the rack 13 6 will result in rotary motion of the gear 130 and the appended cylindrical section 138, and in addition thereto, in rotary motion of the plunger 128 within the bore 126.

As heretofore indicated the shaft 30* is geared to the drive shaft 28 of the engine 20 and driven thereby (FIG. 1). The driven shaft 30 is journale-d into a pair of bearings 146 on both sides of the cam 32 and integral with the housing 122 (FIG. 8). The cam 32 bears against a follower 148 consisting of a globular bearing 150 which rides on the periphery of the cam 32, and extends in ball and socket connection into a cup-shaped member 152 having a collar section 154 adapted to receive the end of the plunger 128 and fastened securely thereto as by set screw 156. The delivery end of the plunger 128 contains a longitudinal channel-shaped slot 158 communicating with a helical strip 160' which winds around the semiperiphery of the plunger 128 so that passage is provided from a fluid chamber 162 through the slot 158 and along a helical offset 161. An inlet port 164 introduces the hydraulic oil from the reservoir 34 into the chamber 162 through an overflow chamber 166 and an inlet passage 168 extending radially through the casing 124 (FIGS. 8 and 9'). A spill port 170 is diametrically disposed from the inlet passage 1'68 and also leads radially through the casing 124 into the overflow chamber 166. However, the spill port 170 is positioned to inevitably coincide with the helical oifset 1 61 as the plunger 128 advances through the fluid chamber 162 during itsdelivery stroke. It is to be noted that the point at which the spill port 170 coincides with the helical offset 161 during the delivery stroke, to vent the hydraulic oil pressure into the overflow chamber 166, may be advanced or retarded by rotating the plunger 128 inside the casing 124-. A return spring 172 compressed between a retainer 174 on the outer housing 122 and a second retainer 176 attached to the collar section 148 provides the motive power for the return stroke of the plunger 128. A check and reverse check valve assembly generally designated by the numeral shaft 116 back towards the chamber 106 thereby tlineally 178 is provided to receive the hydraulic oil from the chamber 162 before distribution to the hydraulic ram mechanism 38 and comprises. an annular inner casing 181) about which a cylindrical outer section 182 is threaded. A bore is provided within the inner casing wherein a valve 184 is adapted to ride responsive to the pressure from the chamber 162. The valve 184 in turn contains a bored chamber 186 from which passageways 188 lead to a groove 190 about the outer periphery of the e7 valve 134 adapted to communicate with a spring chamber 192 and an outlet port 194-. The spring chamber 192 contains a coil spring 196 compressed between the top of the outer casing 1-82 and the valve 184 so as to continuously bias the valve seattoward closed position, which occurs when an inclined surface 1% is in contact with the top of the inner casing 180'. Since the outer casing 182 is threadably engaged to the inner casing 180, an adjustment of the compression in the coil spring 196 is thus provided and may be set as desired depending upon the number of threads in engagement. The chamber 1136 also contains a coil spring 264 bearing on a valve 292 and a threaded plug 264 to form a reverse check valve assembly 286 closing off linear communicating means between the chamber 16-2 and the outlet port 194- by way of a passageway 2% through the plug 204 and a co-linear passageway 21% in the valve 184 to the outlet port 194. Thus it will be apparent that in operation of the unit, as the cam 32 rotates on the driven shaft 3%} and bears against the follower 148 which in turn is attached to the plunger 128, the plunger will be forced upward through the fluid chamber 162 wherein hydraulic oil has been introduced from the reservoir 34 through a conduit 212 (FIG. 1) and the inlet port 164. As the plunger rises the spill port 176" is covered thereby. At this point the plunger 128 starts its delivery stroke and pressure in the system is built up and the valve 184 is forced upward against the force of the coil spring 196, so that hydraulic oil under pressure is expelled from the chamber 162 through the linear passageway 268 into the chamber 186 and thence through the passageway 188 and the groove 1% into the spring chamber 192, and delivered by way of the outlet port 194 and a conduit 214 (FIG. 1) to the pressure relief valve 36, prior to delivery to the hydraulic ram mechanism 38.

As the pump plunger 128 continues its upward motion, the helical offset 161 thereon will come to coincide with the spill port 17%} and the hydraulic oil under pressure will be vented therethrough from the chamber 162 by way of the longitudinal groove 158 and the helical offset into the overflow chamber 166 so that at said point of coincidence, delivery of hydraulic oil to the hydraulic ram mechanism 38 will cease, and the valve 184 will close under the force of the coil spring 196.

It will now be evident that the amount of hydraulic oil expelled from the chamber 162 for delivery to the hydraulic ram mechanism 38 will depend upon the point during the delivery stroke of the plunger 128 at which the helical ofiset 161 coincides with the spill port 170, and it will further be evident that the occurrence of said point may be hastened or retarded by the rotation of the plunger 123 in the casing 124 by the rack and gear assembly 42. i

It will be noted that since the top of the plunger 128 always covers the spill port 176 at the same time, delivery of hydraulic pressure to the hydraulic ram mechanism 38 will always begin at the same time, but the point at which delivery ceases varies with the rotated position of the plunger.

The reverse check valve 2G6 cooperates with the system taining a hydraulic ram 228 adapted to reciprocate therein.

A hydraulic oil cut-elf channel 23th leads from the bottom of the bore 226 around the check valve 226 back into the port 216. In addition thereto a vent channel 2322 leads from an annular groove 234 concentric with the longitudinal bore 226, through the cylindrical casing 215 into the ambient atmosphere.

The hydraulic ram 228 is threadably engaged in an unloader plunger 236 and reciprocates in conjunction therewith. The unloader plunger 236 is a substantially bifurcated channel section, the legs 238 of which extend into a valve seat 240 of the inlet valve 18, and are positioned to seat against a pair of valve channels 242 of the inlet valve. (FIGS. 3 and 4.) Hence the reciprocating motion of the hydraulic ram 228 with the unloader plunger 236 attached thereto opens, and allows closing of the inlet valve 18. A return spring 246 is positioned in a pie-compressed condition between a retaining shoulder 248 in the trough of the unloader plunger 236 and a grooved boss 25% on the valve seat 246, to supply the motive power for the return stroke of the hydraulic ram 223 and the unloader plunger 236.

Referring now to FIGS. 5 and 7, two conditions of particular significance will be noted.- Firstly, the hydraulic ram 228 is concentrically and slidably positioned in the longitudinal bore 228 with an absolute minimum of clearance therebetween (FIG. 5) andsecondly, the bottom of the hydraulic ram culminates in a tapered end section 244 (FIG. 7). The reasons therefore shall appear more fully hereinafter.

It will now be seen that when the hydraulic oil pressure is admitted through the single port 216 from the pump 24 by preventing air bubbles from entering the hydraulic ram mechanism which might otherwise air lock the system and also by insuring that upon seating of the valve 184 against the top of the outer casing 182, hydraulic oil pressure will inevitably be denied to the hydraulic ram mechanism 38.

Operation Hydraulic Ram Mechanism through the check valve 220 into the oil chamber 218, the hydraulic ram 228 will be forced upwards against the force of the return spring 246 together with the attached unloader plunger 236. The legs 238 on the unloader plunger will thus be brought to bear against the valve channels 242. The valve 18 will thus be kept open until the hydraulic oil pressure in the oil chamber 218 exerted against the end section 244 of the hydraulic rarn fails due to hydraulic oil being denied thereto by operation of the pump 24. At this point the compression load on the return spring 246 is suddenly removed, and the spring snaps downward driving the unloader plunger and hydraulic rant on their return stroke with great force and violence.

Means now become operative for closing the inlet valve 18 in a gentle yet positive manner to avoid excessive wear and breakage thereof. As the hydraulic ram 228 advances downward on the return stroke, the check valve 22% immediately closes there-by trapping the remaining hydraulic oil in the oil chamber 218. The hydraulic oil therefore will be expelled through the relatively smaller area of the cut-off channel 230. In this manner the initial deceleration of the hydraulic ram occurs since the hydraulic oil builds up a certain back pressure due to being forced through the smaller cut-oflf channel 230. As

the hydraulic ram 228 further advances on its return 2 stroke (FIGS. 5 and 7) the tapered end section 242 thereof further approaches the cut-oil channel 230 and ultimately covers it completely to close all the last avenue of expulsion for the hydraulic oil. The partially decelerated hydraulic ram now-bears against the remaining oil trapped in the chamber 218 and the check valve 220, so that a dashpot effect occurs and the hydraulic ram 228 is thus brought to a final stop. It will be noted that the hydraulic ram- 228 is fitted into the longitudinal bore 226 to very close tolerance (FIG. 5) so that hydraulic oil cannot seep longitudinally up the bore 226. The tapered end section 244 is designed to further aid thev gradual deceleration process by offering a less resistive surface against the trapped hydraulic oil so that the hydraulic ram.

is brought to a stop in a still more gentle manner.

The annular groove 234 serves the purpose of collecting any small quantity of oil that may penetrate therein from chamber 218, and venting the same to the ambient atmosphere through the vent channel 232. Said oil may also be returned, by suitable conduit, not shown here, to the reservoir 34.

Prior to delivery of hydraulic oil by the pump 24 to the hydraulic ram mechanism 38, the oil is routed through the pressure relief valve 36 (FIG. 6). The pressure relief valve 36 is of a standard type commonly employed in the art and comprises a cylindrical casing 250 having an inlet 252 directly communicable with an outlet 254 by a passageway 256. A tap off 257 to the passageway 256 leads to a valve chamber 258 housing a spring loaded valve assembly 260, adapted to permit ingress of hydraulic oil into the chamber from the tap off 257, but to prevent egress therefrom due to the influence of a coil spring 262 constantly biasing the valve towards closed position. A locking nut 264 threadably mated into the cylindrical casing 250 bears down on the coil spring 262 and is adaptable to fix the compression thereon and consequently to determine the pressure required to open the valve. A relief port 264 communicable with the valve chamber 258 is provided to spill any excess hydraulic oil under pressure passing through the valve assembly 260, back to the hydraulic reservoir 34 by way of a conduit 266 (FIG. 1).

Thus as applied in the herein disclosed system, as the plunger 128 of the pump 24 continues on its delivery stroke, oil pressure is built up in the system. Since the volume of the fluid chamber 162 in the pump is materially greater than the volume of the oil chamber 218 in the hydraulic cam mechanism, the excess hydraulic oil spills through the pressure relief valve assembly 260, and is returned to the hydraulic reservoir 34. Thus pressure is maintained on the hydraulic ram 228 which in turn holds the inlet valve 18 open.

System Operatioru Having thus described the various components of the herein disclosed infinite step control system for a compressor, it will now be seen that in applications such as shown in FIG. 1 where it is desired to maintain a constant pressure source for the operation of pressure tools or instruments (not shown here), the demand is on the re ceiver 44 through a delivery line 268 leading to the con suming source. If the consumption is of such magnitude that it is not immediately replaced by the compressor through the line 270, a lowered pressure value will result in the receiver 44. The lowered pressure thus achieved will pass into the pressure regulator 46 through the conduit 51 and will be introduced into the pressure chamber 50. Now the coil spring 80 will readily overcome the reduced pressure to force the valve stem 66 downward thus increasing the opening of the needle valve assembly 60 thereby introducing more pressure into the second pressure chamber 54. At this point the set pressure of the system may be regulated by the predetermined setting of the coil spring 80. If the compression of the spring is increased a greater pressure will be required to close the needle valve assembly 60. Conversely, if the compression is decreased, a lesser pressure will suffice for the closing operation.

From the chamber 54, the afore-mentioned pressure value is introduced into the conduit 96 and the pressure chamber 106 of the diaphragm motor 40. The diaphragm motor 49 now responds to the new pressure value introduced therein by movement of the motor shaft 116 towards the right (FIG. 1) since the spring 112 is overcome by the pressure introduced into the chamber 106. Thus the rack and gear assembly 42 rotate the pump plunger 128 in a clockwise direction so that the helical groove 160 will coincide with the spill port 170 at an earlier point in the delivery stroke of the plunger (FIG. 8), and in effect shorten the delivery stroke of the pump. In this manner less oil is delivered by the pump to the hydraulic ram mechanism 38. As the ram 228 holds open livered by a gas engine-compressor.

the inlet valve 18 of the compressor 10, and keepsdt open while excess oil is spilled back into the reservoir 34 by the relief valve 36, hydraulic pressure is denied to it sooner so that the inlet valve snaps closed. It will be noted that the amount of time the inlet valve is kept open, is the same as the time required for the delivery stroke of the pump.

At this point, however, because of the timing set between the pump 24 and the compressor 10, the inlet valve 18 will be closing at a relatively early point during the compression stroke of the compressor. Therefore very little fluid will have been expelled from the com-pres sor cylinder 14 back through the valve and the output of the compressor will be greater on that stroke. Thus the compressor will continue to deliver a greater output until the value of the pressure in the receiver is restored.

'It follows that overloading of the receiver cannot occur since this would create a higher pressure signal which following the path above indicated through the components of the system would conversely affect the pump 24 to adjust in a counterclockwise direction thereby lengthening the delivery stroke of the pump plunger 128 and increasing the time thereby that the inlet valve is kept open during the compression stroke of the compressor, to reduce the output thereof. By the system herein disclosed it will be noted that as conditions and requirements vary, the output of the compressor can be regulated from zero percent to one hundred percent of its full capacity.

It will be further noted that when it is desired to operate the compressor at any desired percentage of its capacity, this may simply be accomplished by setting and fixing the plunger 128 of the pump 24 in the proper rotated position. Since the top of the plunger will always cover the spill port at the same time, the inlet valve 18 will always open at thesarne time, and since the offset .on the plunger will always coincide with the spill port at the same time the valve will always close at the same time for any given setting. By proper timing between the pump and the compressor fixing the point during the compression stroke at which closing shall occur,-the capacity of the compressor can be limited to the exact percentage thereof desired.

As an applicable corollary to the infinite step control system, the inlet valve may be made to open before the suction, or intake stroke of the compressor begins; that is to say during the end of the expansion of the fluid in the end clearance of the compressor cylinder. In this manner the efiective intake of fluid into the compressor cylinder is increased. Thus the capacity of the compressor may be increased by taking in more volume of fluid for compression.

The herein disclosed regulatory system may readily be applied to pumping the optimum amount of fluid de- In lieu of the receiver 44, the conduit 51 may be connected to the intake manifold of the engine. The pressure in the intake manifold is a direct indication of the load on the engine, and thus the status of the manifold pressure is introduced into the pressure regulator 46. The pressure regulator 46 may be set by means of the threaded shaft 86 to correspond to the intake manifold pressure at full load of the engine. Thus if the intake manifold pressure rises above its full load pressure, or falls therebelow, due to change in compressor suction and discharge pressures, the pressure output from the regulator 46 will correspondingly rise or fall to operate the diaphragm motor accordingly. The diaphragm motor in turn changes the position of the rack and gear assembly 42 accordingly, to change the output of the pump, thus changing the timing of the compressor inlet valve closing, and consequently the amount of fluid pumped by the attached compressor cylinders.

The operation of the herein disclosed device for closing the inlet valve 18 of the compressor 10 in a gentle yet positive manner is graphically illustrated by FIG. 13. When spillout of the hydraulic oil occurs in pump 24 through the spill port 170, the pressure againstthe ram 228 is suddenly removed. The compressed return spring 245 immediately snaps back driving the unloader plunger 236 and the hydraulic ram 223 before it to allow the valve to close. Normally, without the hydraulic cushioning means herein provided the velocity of said components would increase as the travel thereof increases, along a line 272, so that at the instant of closing contact the closing velocity would attain a high value indicated by a point B.

In accordance with the practice of this invention the velocity of the inlet valve during the travel thereof to closed position is illustrated by a line 274 and is traceable as follows. At the instant that pressure is denied the hydraulic ram 228, the return spring 246- rapidly drives the unloader plunger 236 at normal closing velocity toward point C. Hence as the hydraulic oil is expelled through the cut-off channel 230 at a relatively slower rate, a certain back pressure builds up to decelerate the ram 228. As the ram continues its downward movement,

it covers and seals off the cut-off channel 230 at approximately a point D. Thence the ram is more rapidly decelerated by the pressure of the oil trapped in the chamher 218 against the tapered end 244 and is brought to a gentle yet positive stop at the closing point B. It will be noted that the inlet valve is thus spared the stress and fatigue that it would otherwise have to endure by closing at a rapid velocity measurable by the distance between the points B and E.

It will thus be seen that there are provided novel means for controlling the output of a constant speed compressor, and for closing the inlet valve thereof in a gentle yet positive manner, in which the several objects of this invention are achieved, and which are well adapted to meet the conditions of practical use.

It will be noted that the elimination of clearance pockets as a control media effects a substantial economy in material and permits a smaller and more compact unit for any comparable capacity, while increasing the regulation thereof.

Although only one embodiment of the invention has been illustrated and described, it will readily be apparent to those skilled in the art that changes in form and modifications may be made without departing from the spirit and the scope of the invention.

I claim:

1. In a reciprocating compressor having a piston reciprocating in a compression chamber and an unloading valve which is movable to an open position to prevent said piston from compressing gas in said chamber, a system for regulating the compressed gas output from said chamber by holding said unloading valve open during a variable portion of the compression stroke of the piston, said system comprising: a hydraulic pump; means drivingsaid pump causing it to pump hydraulic fluid pressure during the compression stroke of the compressorv allow said unloading valve to close upon the collapse of hydraulic fluid pressure on said ram; control means for controlling the hydraulic fluid pressure supplied to said ram by said pump during the compression stroke of said compressor piston to vary the period of duration of said hydraulic fluid pressure relative to the duration of said compression stroke for controlling the output of said compressor; and check valve means interconnected be tween said pump and hydraulic ram and operative to allow hydraulic fluid to flow freely from said pump to said ram and to restrict the reverse flow of hydraulic fluid from said ram to said pump to prevent the hydraulic fluid pressure on said ram from collapsing quickly while allowing it to decrease slowly whereby said ram prevents said unloading valve from slamming closed at the release of hydraulic pressure by said control means.

2. The system of claim 1 wherein said check valve means includes: a check valve which is opened by the flow of hydraulic fluid from the pump to the ram and closes upon the reverse flow of fluid from the ram to the fluid dashpot effective-to slow down the movement of said hydraulic ram. during the closing of said unloading valve.

References Cited in the file of this patent UNITED STATES PATENTS Donaldson May 4,

Claims (1)

1. IN A RECIPROCATING COMPRESSOR HAVING A PISTON RECIPROCATING IN A COMPRESSION CHAMBER AND AN UNLOADING VALVE WHICH IS MOVABLE TO AN OPEN POSITION TO PREVENT SAID PISTON FROM COMPRESSING GAS IN SAID CHAMBER, A SYSTEM FOR REGULATING THE COMPRESSED GAS OUTPUT FROM SAID CHAMBER BY HOLDING SAID UNLOADING VALVE OPEN DURING A VARIABLE PORTION OF THE COMPRESSION STROKE OF THE PISTON, SAID SYSTEM COMPRISING: A HYDRAULIC PUMP; MEANS DRIVING SAID PUMP CAUSING IT TO PUMP HYDRAULIC FLUID PRESSURE DURING THE COMPRESSION STROKE OF THE COMPRESSOR PISTON; A HYDRAULIC RAM CONNECTED TO SAID PUMP TO RECEIVE THE HYDRAULIC FLUID PRESSURE PRODUCED BY SAID PUMP, SAID HYDRAULIC RAM BEING OPERATIVE, IN RESPONSE TO THE HYDRAULIC PRESSURE RECEIVED FROM SAID PUMP, TO MOVE TO A POSITION WHEREIN IT HOLDS SAID UNLOADING VALVE IN AN OPEN POSITION DURING A PORTION OF THE COMPRESSION STROKE OF SAID COMPRESSOR, TO UNLOAD SAID COMPRESSOR, AND TO ALLOW SAID UNLOADING VALVE TO CLOSE UPON THE COLLAPSE OF HYDRAULIC FLUID PRESSURE ON SAID RAM; CONTROL MEANS FOR CONTROLLING THE HYDRAULIC FLUID PRESSURE SUPPLIED TO SAID RAM BY SAID PUMP DURING THE COMPRESSION STROKE OF SAID COMPRESSOR PISTON TO VARY THE PERIOD OF DURATION OF SAID HYDRAULIC FLUID PRESSURE RELATIVE TO THE DURATION OF SAID COMPRESSION STROKE FOR CONTROLLING THE OUTPUT OF SAID COMPRESSOR; AND CHECK VALVE MEANS INTERCONNECTED BETWEEN SAID PUMP AND HYDRAULIC RAM AND OPERATIVE TO ALLOW HYDRAULIC FLUID TO FLOW FREELY FROM SAID PUMP TO SAID RAM AND TO RESTRICT THE REVERSE FLOW OF HYDRAULIC FLUID FROM SAID RAM TO SAID PUMP TO PREVENT THE HYDRAULIC FLUID PRESSURE ON SAID RAM FROM COLLAPSING QUICKLY WHILE ALLOWING IT TO DECREASE SLOWLY WHEREBY SAID RAM PREVENTS SAID UNLOADING VALVE FROM SLAMMING CLOSED AT THE RELEASE OF HYDRAULIC PRESSURE BY SAID CONTROL MEANS.

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