Lubrication system of reciprocating piston type compressor for engine supercharging

Abstract: To obtain accurate lubrication of a reciprocating piston type compressor for supercharging without affecting the operation of an engine oil pump, using engine lubricating oil. A lubricating oilway 89 for the compressor is branched off from a lubricating oilway 87 for an engine E fitted with an oil pump P 1 . A downstream end of this branched oilway 89 opens into an annular groove 77 connected to an operating chamber 37 of the inner surfaces of a pump cylinder body 20 of the reciprocating piston type compressor C for supercharging. A second oil pump P 2 for supplying some of the lubricating oil flowing in the lubricating oilway 87 for the engine to the pump cylinder body 20 side, and having a smaller capacity than the oil pump P 1 , is fitted into the lubricating oilway 89 for the compressor.

Description: Detailed Description Of The Invention Technical Field The present invention relates to a lubrication system of a reciprocating piston type compressor for engine supercharging, and particularly to a lubrication system of a reciprocating piston type compressor for engine supercharging provided with a pump cylinder body, pump pistons, fitted into a cylinder hole of the pump cylinder body at both ends of the cylinder hole, defining first and second pump chambers communicating with a supercharging port of an engine, and drive means, arranged in a drive chamber formed in a central part of the pump piston for supplying reciprocating movement to the pump piston by driving using a drive shaft of the engine. Related Art Up to now, as a lubrication system for an engine supercharging compressor, it has been known to branch a compressor lubrication oilway from a pressurized oilway between an oil pump and a lubrication section, inside an engine lubrication oilway sequentially communicating with the oil pump, lubrication section and oil pan of an engine, to open the compressor lubrication oilway onto a sliding surface of the compressor, and to supply engine lubrication oil to the compressor using discharge pressure of the oil pump (for example, to Japanese patent Publication No. Sho. 56-10451). Problems To Be Solved By The Invention Conventionally, the capacity of the engine oil pump has primarily been determined depending on the various parts of the engine to be lubricated, which means that if engine lubrication oil is supplied to the compressor using this oil pump, as has conventionally been the case, the supply capacity of the pump was determined by the operating conditions of the compressor, and it was difficult to carry out fine control. The present invention has been conceived in order to solve the above described problems, and an object of the invention is provide a lubrication system for an engine supercharging compressor in which the compressor is precisely lubricated according to its operating conditions, even while using the engine lubricating oil. Means Of Solving The Problems in order to achieve the above described object, a first aspect of the present invention is directed to a lubrication system of a reciprocating piston type compressor for engine supercharging, comprising a pump cylinder body, pump pistons, fitted into a cylinder hole of the pump cylinder body at both ends of the cylinder hole, defining first and second pump chambers communicating with a supercharging port of an engine, and drive means, arranged in an operation chamber formed in a central part of the pump piston for supplying reciprocating movement to the pump piston by driving using a drive shaft of the engine, wherein a compressor lubrication oilway is branched off from an engine lubrication oilway sequentially communicating with an oil pump, a lubrication section and an oil pan of the engine, a downstream end of this compressor lubrication oilway opens out into a section, at an inner surface of the pump cylinder body, communicating with the drive chamber, and a second pump having a smaller capacity than the oil pump, for supplying some of the lubricating oil flowing in the engine lubrication oilway to the pump cylinder body side, is fitted into the compressor lubrication oilway. According to this first aspect, by controlling the second oil pump there is no affect on the operation of the engine oil pump, and an appropriate amount of lubricating oil can be drawn out from and engine lubricating oilway depending of the operating conditions of the reciprocating piston type compressor and supplied to the compressor. As well as lubricating reciprocal sliding surfaces of the pump cylinder body and the pump piston, the lubricating oil supplied to the compressor can lubricate various parts of the drive means by penetrating even during operation. In addition to the above described aspect, a second aspect of the present invention further comprises an annular groove facing the operation chamber being formed on an inner surface of the pump cylinder body, and the downstream end of this compressor lubrication oilway opening out into this annular groove. According to the second aspect, an oil sump is formed between the annular groove and the outer surfaces of the pump piston and lubricating oil is retained, which means that lubrication of reciprocally sliding surfaces of the pump cylinder body and the pump piston can be carried out effectively. |Effects Of The Invention| According to a first aspect of the present invention there is provided a lubrication system of a reciprocating piston type compressor for engine supercharging, comprising a pump cylinder body, pump pistons, fitted into a cylinder hole of the pump cylinder body at both ends of the cylinder hole, for defining first and second pump chambers communicating with a supercharging port of an engine, and drive means, arranged in a drive chamber formed in a central part of the pump piston for supplying reciprocating movement to the pump piston by driving using a drive shaft of the engine, wherein a compressor lubrication oilway is branched off from an engine lubrication oilway sequentially communicating with an oil pump, a lubrication section and an oil pan of the engine, a downstream end of this compressor lubrication oilway opens out into a section, at an inner surface of the pump cylinder body, communicating with the operation chamber, and a second pump having a smaller capacity than the oil pump, for supplying some of the lubricating oil flowing in the engine lubrication oilway to the pump cylinder body side, is fitted into the compressor lubrication oilway. This means that by controlling the second oil pump there is no affect on the operation of the engine oil pump, and an appropriate amount of lubricating oil is drawn out from and engine lubricating oilway depending of the operating conditions of the reciprocating piston type compressor and supplied to the compressor, making it possible to lubricate reciprocal sliding surfaces of the pump cylinder body and the pump piston, and to supply an appropriate amount of lubricating oil to drive means for the drive chamber. According to a second aspect of the present invention, there is provided an annular groove facing the operation chamber formed on an inner surface of the pump cylinder body and the downstream end of this compressor lubrication oilway opens out into this annular groove. This means that an oil sump is formed in the annular groove between the outer surfaces of the pump piston and lubricating oil is held in this sump, so it is possible to effectively lubricate reciprocal sliding surfaces of the pump cylinder body and the pump piston. Brief Description Of The Drawings Fig. 1 is a cross-sectional side view of a supercharged engine, for use with a motorcycle, of an embodiment of the present invention. Fig. 2 is a cross section along line 2 - 2 in Fig. 1. Fig. 3 is a cross section along line 3-3 3 in Fig. 1. Fig. 4 is an enlarged cross sectional side view of the reciprocating piston type compressor in Fig. 1. Fig. 5 is a cross section along line 5 - 5 in Fig. 4. Fig. 6 is a cross section along line 6 - 6 in Fig. 5. Fig. 7 is a drawing in the direction of arrow 7 in Fig. 4. Fig. 8 is a plan view of disassembled essential parts of the above described compressor. Fig. 9 is a cross sectional view of the disassembled essential parts of the above described compressor. Fig. 10 is a chart showing operation open and close timing of the inlet valve, exhaust valve and supercharging valve of the above described engine, and operation timing of pump piston of the above described compressor. Fig. 11 is a chart showing variations in rotation angle of a pump piston side needle bearing in the above described compressor. Fig. 12 is a cross sectional view of a plunger pump for supplying lubricating oil to the compressor. Fig. 13 is a cross sectional view along line 13 - 13 in Fig. 12. Fig. 14 is a cross sectional view showing a modified example of the drive section of the plunger pump. Embodiments Of The Invention Embodiments of the present invention will now be described based on practical examples shown in the attached drawings. In Fig. 1 to Fig. 3, reference numeral E represents a supercharged engine mounted in a motor cycle as a power source. An engine body 1 of this engine E is comprised of a cylinder head 1b bolted to an upper surface of a cylinder block 1a, and a combustion chamber 3 facing a head of a piston 2 contained in a cylinder bore 8 of the cylinder block 1a, as well as an intake port 4, exhaust port 5 and supercharging port 6 respectively opening out into the combustion chamber 3, are formed in this cylinder head 1b. The diameters of these ports are set such that the intake port 4 is the largest, the supercharging port is the next largest and the exhaust port is the smallest. The inner ends of the intake port 4 and the exhaust port 5 are arranged along a straight line X orthogonal to the axis Y of the cylinder bore 8 and are lined up either side of the axis Y. Also, the inside of the supercharging port 6 partially interjects between the insides of the intake port 4 and the exhaust port 5. The intake port 4 and the exhaust port 5 communicate with an intake manifold and an exhaust manifold (neither of which are shown in the drawings), as in a normal engine, but the supercharging port 6 communicates with a discharge duct 57 of a reciprocating piston type compressor C arranged adjoining an outer end of a supercharging cam 11c. An intake valve 7i, exhaust valve 7e and supercharging valve 7c respectively opening and closing the insides of the intake port 4, the exhaust port 5 and the supercharging port 6, a spark plug 10 for igniting a fuel/air mixture introduced into the combustion chamber 3, and a single valve gear cam shaft 11 for opening and closing the three valves 7i, 7e and 7c are attached to the cylinder head 1b. An electrode 10a of the spark plug 10 is arranged so as to protrude between the inside of the intake port 4 and the exhaust port 5 at the opposite side to the inside of the supercharging port 6. Also, the valve gear cam shaft 11 is arranged along the straight line X, and the intake valve 7i and the exhaust valve 7e are arranged either side of the axis Y and aligned so as to form a V-shape along the straight line X. This makes it possible to arrange the supercharging cam shaft 11c between the intake valve 7i and the exhaust valve 7e, thus enabling reduction in the size of the engine E. The supercharging valve 7c is arranged in parallel to the axis Y. Valve springs 12i, 12e and 12c are respectively attached to each of the valves 7i, 7e and 7c to exert force in a direction of closing them. The valve gear cam shaft 11 is supported on the cylinder head 1b via a pair of left and right ball bearings 15 and 16. A tapered intake cam 11i and a tapered exhaust cam 11e, directly engaging with each valve head of the intake calve 7i and the exhaust valve 7e and arranged between the two bearings 15 and 16, and a supercharging cam shaft 11c engaging with the valve head of the supercharging valve 7c via a rocker arm 19 axially supported on the cylinder head 1b, and arranged between the cams 11i and 11e, are formed on the valve gear cam shaft 11. The inside of the intake port 4 and the exhaust port 5 are also arranged so as to completely face the cylinder bore 8, and the inside of the supercharging port 6 is arranged so that part of it protrudes to the outside of the cylinder bore 8. In other words, part of the combustion chamber 3 is caused to protrude outside the cylinder bore 8, and the supercharging valve 7c is arranged so that part of the supercharging port 6 is hollowed out at the protruding part of the combustion chamber 3. Accompanying this, part of the supercharging valve 7c also protrudes outside of the cylinder bore 8, and a gap g that is slightly larger than the opening and closing stoke of the supercharging valve 7c is provided between the protruding part and an upper surface of the cylinder block 1a. The opening and closing stroke of the supercharging valve 7c is much smaller than the opening and closing stroke of the intake valve 7i and exhaust valve 7e. A first squish area 13 of the combustion chamber 3 is defined between the supercharging valve 7c and one side of a head surface of the piston 2 flattening out at the upper limit of travel of the piston 2, and a second squish area 14 of the combustion chamber 3 is defined between the other side of the head surface of the piston 2, also flattening out at the upper limit of travel of the piston 2. The combustion chamber 3 of the cylinder head 1b has a deepest section 3a offset to the sparkplug 10 side from the axis Y, and a concave section 2a corresponding to this deepest section 3a is formed in the head surface of the piston 2. A driven sprocket 18 driven from a crank shaft (not shown) connected to the piston 2 via a chain 17 is fastened to an end of the valve gear cam shaft 11 that projects outside the left bearing 15. Further, if the valve gear cam shaft 11 is driven from the crank shaft via the chain 17, the intake exhaust and supercharging valves 7I, 7e and 7c respectively open and close at the timing shown in Fig. 10 by cooperation of the valves 11i, 11e and 11c for intake exhaust and supercharging, as well as the valve springs 12, 13 and 14. Accordingly, spanning from the intake stroke to the compression stroke, the supercharging valve 7c is only open for the fixed time period from immediately before the intake valve 7i closes until after the valve is closed. During this valve opening time, namely from after the intake stroke until the compression stroke has commenced, high pressure air from the compressor C is supercharged from the supercharging port 6 into the combustion chamber 3. As a result, charging efficiency is increased, and it is possible for the engine E to exhibit high output power. Particularly, part of the inside of the supercharging port 6 is arranged protruding from the cylinder bore 8, and so it becomes possible to sufficiently enlarge the diameter of the supercharging port 6, sufficient supercharging is obtained without completely sacrificing the overall diameter of the intake port 4 and the exhaust port 5, and it is possible to effectively increase charging efficiency. Also, since there is no need to specially increase the diameter of the cylinder bore 8, it is not easy for knocking to occur. In this case, part of the supercharging valve 7c for opening and closing the inside of the supercharging port 6 also projects from the cylinder bore 8 and opposes the upper surface of the cylinder block 1a, and the gap g slightly larger than the opening and closing stroke of the supercharging valve 7c is provided between these opposing surfaces. This means that the supercharging valve 7c can be opened and closed without interfering with the cylinder block 1a. Also, since the opening and closing stroke of the supercharging valve 7c is much smaller than that of the intake valve 7l and the exhaust valve 7e, reduction in compression ratio accompanying formation of the gap g is comparatively small. The inner ends of the intake port 4 and the exhaust port 5 opening into the combustion chamber 3 are arranged along a straight line X orthogonal to the axis Y of the cylinder bore 8, and either side of the axis Y, and the inner end of the supercharging port 6 is arranged so as to partially project between the inner ends of the intake port 4 and exhaust port 5. Therefore, part of the inner end of the supercharging port 6 is caused to project outside the cylinder bore 8, it is possible to sufficiently increase the diameter without each of the ports 4, 5 and 6 interfering with each other, and further improvements to charging efficiency and exhaust efficiency can be obtained. On the other hand, the electrode 10a of the spark plug 10 is arranged projecting between the inner ends of the intake port 4 and the exhaust port 5 at the opposite side to the supercharging port 6, which means that the electrode 10a can be allowed to approach the central section of the combustion chamber 3 without obstructing the intake port 4, exhaust port 5 and supercharging port 6, with the result that the propagation time taken for a flame occurring at the time of ignition to reach all edges of the combustion chamber is shortened considerably, which can contribute to the prevention of knocking. Using the supercharging valve 7c, a larger first squish area 13 of the combustion chamber 3 is defined between the supercharging valve 6 and one side of the head surface of the piston 2, while a second squish area 14 of the combustion chamber 3 is defined between the cylinder head 1b and the other side of the head surface of the piston 2. This means that in the period of time after the compression stroke the mixture in the combustion chamber 3 is strongly compressed by the first and second squish areas 13 and 14 and quickly forced to the side of the deepest section 3a of the combustion chamber 3, with the result that the mixture in the combustion chamber 3 is well stirred, and at the time of ignition the flame propagation time is improved contributing to prevention of knocking and rarefied combustion. The deepest part 3a of the combustion chamber 3 is located offset to the spark plug 10 side from the axis Y of the cylinder bore 8 and the concave section 2a is formed in the head surface of the piston 2 corresponding to the deepest section 3a, which means that at the time of ignition it is easy to generate a flame in the deepest section 3a and the hollow section 2a, and this flame spreads smoothly from the deepest section 3a and the hollow section 2a to edges including the first and second squish regions 13 and 14, and it is possible to obtain a favorable combustion state of the mixture. Next, description will be given of a reciprocating piston type compressor C, using Fig. 4 to Fig. 7. A reciprocating piston type compressor C comprises a pump cylinder body 20 having bearing bosses 21 and 22 projecting beyond both left and right outer side surfaces, a pump piston 25 slidably fitted into a cylinder hole 24 of the pump cylinder body 20, and a pump crankshaft 26 for driving this pump piston 25. The pump cylinder body 20 has the left bearing boss 21 fitted into an attachment hole27 on a right side wall of the cylinder head 1b and fastened with bolts 28 (refer to Fig. 7). The pump piston 25 is not provided with a piston ring, but slides directly inside the cylinder hole 24 of the pump cylinder body 20, and lubricating grease is coated onto the sliding surfaces of the pump piston 25 and the cylinder hole 24. There are bearing holes 22a and 23a reaching as far as the inner surface of the pump cylinder body 20 in the bearing bosses 21 and 22, a pump crankshaft 26 is supported by ball bearings 29 and 30 fitted into these bearing holes 21a and 22a, and one end of the pump crankshaft 26 is connected to the valve gear cam shaft 11 via splines 31. An oil seal 32 for tightening around the outer edge of the pump crankshaft 26 outside the bearing 29 is fitted into to the left bearing hole 21a. This oil seal 32 is constructed as a high pressure type to the extent that it can withstand the valve opening pressure of a relief valve 75 that will be described later. A seal plug 32 adjoining the outside surface of the bearing 30 is fitted into the bearing hole 22a, and a cap 34 for covering this seal plug 33 is screwed to the outer edge of the bearing boss 22. The two ends of the cylinder hole 24 of the pump cylinder body 20 are closed off by first and second pump cylinder heads 23 1 and 23 2 forming a pair, and first and second piston heads 25 1 and 25 2 defining first and second pump chambers 36 1 and 36 2 between these pump cylinder heads 23 1 and 23 2 are formed on either end of the pump piston 25. A circular operating chamber 37 penetrating between the two piston heads 25 1 and 25 2 and biased to the side of the second piston head 25 2 and a piston pin hole 38 penetrating through the first piston head 25 1 in a sideways direction for supporting the piston pin 39 are formed in the pump piston 25, and a crank pin 26a of the pump crankshaft 26, and a connecting rod 40 for connecting this crank pin 26a to the piston pin 39, are housed in the operating chamber 37. The connecting rod 40 has a first bearing hole 40a in an end at the cranks pin 26a side and a second bearing hole 40b in an end at the piston pin 39 side, and the crank pin 26a and the piston pin 39 are respectively supported by first and second needle bearings 41 and 42 fitted into these bearing holes 40a and 40b. An oil sump 46 is formed in the connecting rod 40, and oil holes 47 and 48 respectively leading to the two bearing holes 40a and 40b are cut through this oil sump 46. In order to simplify the manufacture of the pump piston 25, the pump piston is split up into two piston half bodies 25a and 25b between the operating chamber 37 and the piston pin hole 38, and these two piston half bodies are fastened together using a plurality of bolts 49. As shown in Fig. 4 to Fig. 6, engagement holes 50 in opposite surfaces of the pump cylinder body 20, annular discharge chambers 51 having a smaller diameter than the engagement holes 50, and cylindrical intake chambers 52 surrounding the discharge chambers 51 are respectively provided in the first and second pump cylinder heads 23 1 and 23 2 , and with these engagement holes engaged with both outer ends of the pump cylinder body 20 the two pump cylinder heads 23 1 and 23 2 are integrally fastened together using a plurality of bolts 53 and nuts 54. Also, a first connecting pipe 55 1 for connecting between these two intake chambers 52, 52, and a second connecting pipe for connecting between the two discharge chambers 51, 51, are respectively attached to the two pump cylinder heads 23 1 and 23 2 , and an intake duct 56 for connecting the intake chamber 52 to a central portion of an intake manifold (not shown) of the engine E, and a discharge duct 57 for connecting the discharge chamber 51 to the supercharging port 6 of the engine E are connected to the second pump cylinder head 23 2 . Valve systems 58 for the engagement holes 50 are provided between the pump cylinder body 20 and each of the pump cylinder heads 23 1 and 23 2 , as described in the following. As shown in Fig. 8 and Fig. 9, a valve system 58 is composed of an annular dorsal plate 60, a thin intake valve plate 61, a valve seat plate 62 and a thin discharge valve plate 63, stacked in that order on top of each other. These structural members 60, 61, 62 and 63 are formed circularly having substantially the same outer diameter as the end sections of the pump cylinder body 20. These valve systems 58 have the dorsal plate 60 located on the end surface side of the pump cylinder body 20, are fitted inside the engagement holes 50 of the pump cylinder heads 23 1 and 23 2 corresponding to the end sections of the pump cylinder body 20 and are held between the pump cylinder body 20 and each of the pump cylinder heads 23 1 and 23 2 using the bonding force connecting between the first and second pump cylinder heads 23 1 and 23 2 while inserting them into the pump cylinder body 20 provided by the bolts 53 and nuts 54, as described above. In this case, a first knock pin 65 1 is engaged in first locating holes 64 1 provided in each of the pump cylinder heads 23 1 and 23 2 ; the discharge valve plate 63 and the valve seat plate 62, while a second knock pin 65 2 is engaged in second locating holes 64 2 provided in the valve seat plate 62, the intake valve plate 61 and the dorsal plate 60. Four groups of three intake holes 67 are cut close to the center of the valve seat plate 62 and spaced 90°apart in the circumferential direction, and two groups of 7 discharge holes 68 are cut close to the outside of the valve seat plate 62 spaced 180° apart from one another in the circumferential direction. Four intake reed valves 61a corresponding to the four sets of intake holes 67, and two elongated arc shaped holes 69 respectively encircling the two sets of discharge holes 68 so as not to block them off, are provided in the intake valve sheet 61. Each intake reed valve 61a has a base end extremely close to the outer edge of the intake valve plate 61 and a tip end extremely close to the center of the intake valve plate 61, and is formed by cutting slits in the intake valve plate 61 along its perimeter so as to extend in a radial direction of the intake valve plate 61. Notch-like regulating sections 60a corresponding to the base ends of each of the intake reed valves 61a are provided on the inside of the dorsal plate 60, and bending fulcrums of the intake reed valves 61a are regulated by these regulating sections 60a. By forming the regulating sections 60a as these notches, the bending length of the intake reed valves 61a can be made extremely long without interfering with opening edges of the cylinder hole 24 of the pump cylinder body 20. If it is desired to make this bending length short, the regulating sections 60a can be made to jut out. Two discharge reed valves 63a corresponding to the two groups of discharge holes 68, and a large circular hole 70 encircling the four sets of intake holes 67 so as not to block them off, are provided in the discharge valve plate 63. The discharge reed valves 63a are formed by cutting slits in the discharge valve plate 63 along its perimeter. An annular partition 62a engaging with inner surfaces of the intake chambers 52 of corresponding pump cylinder heads 23 1 and 23 2 is formed on the upper surface of the valve seat plate 62 passing through the circular hole 70, and this annular partition divides the intake chamber 52 from the discharge chamber 51. As shown in Fig. 5, a valve attachment hole 71 opening into an outer side surface of the pump cylinder body 20 , and a relief hole 72 penetrating through a bottom wall of the valve attachment hole 71 are provided in a side wall of the pump cylinder body 20, and an annular recess 77 connecting the relief hole 72 with the operation chamber 37 is formed at part of the inner surface of the pump cylinder body 20 opposite to the pump piston 25. A valve housing 73 fits over the outside of the first connecting pipe 55 1 , and a plurality of communication holes 74 for connecting the inside of the valve housing 73 with the inside of the connecting pipe 551 are formed in the peripheral walls of the connecting pipe 55 1 . A relief valve 75 for opening and closing the relief holes 72, and a valve spring for urging the relief valve 75 in a closing direction against a specified set load, are contained in the valve housing 73. In Fig. 4, reference numeral 80 represents inspection holes provided in each of the pump cylinder heads 23 1 and 23 2 , and reaching to the intake chamber 52, and they are normally closed off by bolts 81. During running of the engine E, if the pump crankshaft 26 of the compressor C is driven by the valve gear cam shaft 11, the pump piston 25 is forcedly supplied with reciprocating movement through the connecting rod 40 and the first and second pump chambers 36 1 and 36 2 alternate between decompression and compression. At the time of decompression of the first pump chamber 36 1 , due to the fact that the discharge holes 68 are closed off by the discharge reed valve 63a and the intake holes 67 are open by the intake reed valve 61a, air inside the intake manifold (not shown) of the engine E is inducted from the intake duct 56 passing sequentially through the first connecting pipe 55 1 , the intake chamber 52 and the intake holes 67 to the pump chamber 36 1 . Also, at the time of compression of the pump chamber 36 1 , since the intake holes 67 are closed off by the intake reed valve 61a and the discharge holes 658 are opened by the discharge reed valve 63a, compressed air in the pump chamber 36 1 is supplied from the discharge holes 68, passing sequentially through the discharge chamber 51, the second connecting pipe 55 2 and the discharge duct 57, to the supercharging port 6 of the engine E. At the time of decompression of the second pump chamber 36 2 , similarly to the case for the first pump chamber 361, due to the fact that the discharge reed valve 63a is closed and the intake reed valve 61a is open, air inside the intake manifold of the engine E is inducted from the intake duct 56 without passing through the first connecting pipe 55 1 , to the intake chamber 52, the intake holes 67 and the pump chamber 36 2 . Also, at the time of compression of the pump chamber 36 2 , similarly to the case for the first pump chamber 361, since the intake reed valve 61a is closed and the discharge reed valve 63a is open, compressed air in the pump chamber 36 2 is discharged from the discharge holes 68 to the discharge chamber 51 and the discharge duct 57, without passing through the second connecting pipe 55 2 and supplied to the supercharging port 6 of the engine E. In the compressor C, the first and second pump chambers 36 1 and 36 2 are alternately caused to operate by a single pump piston 25, which means that the size of each part of the pump piston 25 is made small for overall amount of flow per unit time, and it is possible to realize significant reduction in size of the compressor C. Also, since the pump crankshaft 26 is supported by a pair of ball bearings 29 and 30 on both side walls of the pump cylinder body 20, it is possible to firmly support the pump crankshaft 26 using the high rigidity pump cylinder body 20. Since the pump piston 25 is driven from the pump crankshaft 26 through the connecting rod 40, variation on the swinging speed of the connecting rod 40 relative to the piston pin 39 during rotation of the pump crankshaft 26 is smooth, and so, as shown in Fig. 11, variations in the rotational speed of a needle bearing 42 supporting the piston pin 39 are also always smooth and it is possible to increase the durability of the needle bearing 42. The pump piston 25 slides directly on the inside of the cylinder hole 24 of the pump cylinder body 20 without the provision of piston rings, which means that power loss due to the sliding resistance of piston rings can be kept small. However, since there are no piston rings, compressed air of each of the pump chambers 36 1 and 36 2 passes through a miniscule gap between the pump piston 25 and the inner surface of the cylinder hole 24 and leakage to the operating chamber 37 is unavoidable, but by causing the operating chamber to be pressurized using this leakage a difference between the pressure at the time of compression of the first and second pump chambers 36 1 and 36 2 and the pressure of the operating chamber 37 is reduced. As a result, the amount of gas leaking to the operating chamber 37 is reduced, and it is possible to increase the operating efficiency of each of the pump chambers 36 1 and 36 2 . In the case that the inside of the operating chamber 37 has been pressurized to equal to or greater than a specified pressure, the relief valve 75 is opened and excess pressure of the operating chamber 37 is released to the low pressure first connecting pipe 55 1 , which means that over-pressurization of the operating chamber 37 is prevented, it is possible to increase the durability of other seal portions other than the oil seal 32, and it is possible to prevent leakage of gas from the operating chamber 37 to the cylinder head 1b of the engine E. Since the gas that has been released to the first connecting pipe 55 1 is again taken in to the first and second pump chambers 36 1 and 36 2 , there is no release of gas to the outside and no waste. Lubrication systems for the engine E and the reciprocating piston type compressor C will now be described with reference to Fig. 1 and Fig. 13. First of all, as in the normal case, the lubrication system of the engine E comprises an oil pan 83 for holding lubricating oil, an oil pump P 1 for sucking up lubricating oil from the oil pan 83 through a strainer 84, a pressurized oilway 87a for guiding lubrication oil that has been discharged by the oil pump P 1 to lubrication sections 86 (a piston, crankshaft, valve mechanism, etc.) inside the engine E, and a low pressure oilway 87b for returning lubricating oil that has completed lubrication of the lubrication sections 86 to the oil pan 83. Lubricating oilways 87 for the engine E are constituted by the pressurized oilway 87a and low pressure oil way 87b. Next description will be given of the lubrication system of the reciprocating piston type compressor C. A bypass oilway 88 for bypassing the lubricating sections 86 is connected to the lubricating oilways 87 for the engine, and branched oilway 89 branched midway along the bypass oilway 88 is connected to the reciprocating piston type compressor C through a plunger pump P 2 (second oil pump). An orifice 90 is provided in an upstream part of the bypass oilway 88, and acts as a regulator so that at least a required mount of lubricating oil flows from the pressurized oilway 87a to the bypass oilway 88. The plunger pump P 2 is comprised of a pump body 92 fastened to an outer side surface of the pump cylinder body 20 of the compressor C, a plunger 91 defining a pump chamber 93 and slidably fitted into a cylinder hole 92a of the pump body 92, a solenoid 95, attached to an upper surface of the pump body 92 for imparting a discharge function on the plunger 91 via a movable core 93 upon excitation, and a return spring 97 for urging the plunger 91 in an intake function direction via a retainer 96. The plunger pump P 2 has a lower capacity than the oil pump P 1 of the engine E. An intake port 98 for connecting the branched oilway 89 to the pump chamber 93, and a discharge port 99 for connecting the pump chamber 93 to an inner surface of the pump cylinder body 20 of the compressor C opposite to the pump piston 25, specifically, the annular groove 77, are provided in the pump body 92, and an intake valve 100 and a discharge valve 101 are respectively installed in the intake port 98 and the discharge port 99. A cover 102 for covering the solenoid 95 is also fitted into the pump body 92, and an adjustment bolt 103 for adjusting the stroke of the movable core 94 is screwed into the cover 102. The branched oilway 89 has a specified volume to allow it to have a sump function, and is provided with a specified difference H between an upper inlet and a lower outlet . A filter 104 is provided at the inlet side of the branched oilway 89, so as to expose a filtration surface to a flow of lubricating oil for the engine E inside the bypass oilway 88. During running of the engine, a lot of the lubricating oil discharged by the oil pump P 1 is supplied to the lubricating sections 86 through the pressurized oilway 87a, but some is made to flow to the bypass oilway 88 while being regulated by the orifice 90. Some of this portion of the lubricating oil flows to the branched oilway 89 side, while after it has passed through the by pass oilway 88 the remainder is recombined in the low pressure oilway 87b with lubricating oil that has finished lubricating the lubricating sections 86 and flows back to the oil pan 83. During this operation, the plunger 91 is carrying out a pumping operation in the plunger pump P 2 due to repeated excitation and demagnetization of the solenoid 95, which means that during the intake stroke lubricating oil is inducted from the branched oilway 89 to the pump chamber 93, and during the discharge stroke where lubricating oil is supplied from the discharge port 99 to the annular groove 77 on the inner surface of the pump cylinder body 20. The annular groove 77 forms an oil sump between the outer surfaces of the pump piston 25, and mutually sliding surfaces of the pump cylinder body 20 and the pump piston 25 are effectively lubricated by the oil retained in this sump. As well as this, lubricating oil that has been supplied to the annular groove 77 also flows in to the pump chamber 37 of the pump piston 2 and is scattered by the reciprocating movement of the pump piston 25 to lubricate the bearings 29 and 30 supporting the pump crankshaft 26, and is then held in the sump 46 of the connecting rod 40, passes through the oil holes 47 and 48 and lubricates the needle bearings 41 and 42 respectively supporting the crank pin 26a and the piston pin 39. These various flows of lubricating oil leak out to the first and second pump chambers 36 1 and 36 2 through a sliding surface gap between the pump cylinder body 20 and the pump piston 25, accompany the reciprocating movement of the pump piston 25, and the oil is used up by being supplied to the engine E together with supercharging air. In order to compensate for the used up amount, a lubricating oil discharge amount of the plunger pump P 2 is adjusted by controlling the number of times the solenoid 95 is energized. Accordingly, while using lubricating oil of the engine E in the lubrication of the compressor C, it is possible to supply an appropriate amount of lubricating oil in small amounts according to the operating conditions to the compressor C, without affecting the operation of the oil pump P 1 of the engine. Lubricating oil is usually flowing in the bypass oilway 88, and since bubble retention does not occur, it is possible to supply oil free from bubbles into the branched oilway 89 through the filter 104. Also, this branched oilway 89 has an oil retaining function and has an inlet facing upwards, which means that since only a little lubricating oil flows in the branched oilway 89, even if bubbles do arise in this oil these bubbles immediately climb the branched oilway 89 and move to the bypass oilway 88, and are discharged to the oil pan 83 together with lubricating oil that passes through the bypass oilway 88. Accordingly, the plunger pump P 2 always supplies an appropriate amount of bubble free lubricating oil from the branched oilway 89 to the compressor C, and this enables lubrication to be carried out accurately. In addition, the filtering surface of the filter 104 is exposed to the flow of engine lubricating oil in the bypass oilway 88 which means that the filtering surface is always cleaned and there is no build up of foreign substances, a so-called self cleaning effect is obtained, and it is possible to preemptively prevent a reduction in the intake amount of the plunger pump P 2 that would be caused by silting up of the filtering . surface. Also, since a specified difference H is provided between the upper inlet and lower outlet of the branched oilway 89, the weight of the lubricating oil inside the branched oilway 89 is applied to the intake port 98 of the compressor C and the plunger pump P 2 can supply lubricating oil at the same time as commencing operation without delaying the compressor C. Fig. 14 shows a modified example where the drive section of the above described plunger pump P 2 adopts a negative pressure drive system in place of a solenoid drive system. In this modified example, an operating piston 105 linked to the upper end of the plunger 91 is slidably engaged with a cylinder 106 of the pump body 92. The inside of this cylinder 106 is divided into an upper atmospheric air chamber 107 and a lower transformation chamber 108 by the operating piston 105, and an extending pathway 109 of the transformation chamber 108 is provided so as to switchably connect alternately to the an atmospheric air pathway 111 or a negative pressure pathway 112 via a magnetic switching valve 110. The atmospheric pathway 111 is open to the atmosphere, while the negative pressure pathway 112 is connected to a source of negative pressure (for example, the inside of the intake manifold of the engine E). A return spring 113 for urging the plunger 91 in the direction of an intake operation via the operating piston 105 is housed in the transformation chamber 108. The remainder of the structure is the same as the above described plunger pump P 2 , and so parts in the drawing corresponding to parts of the plunger pump P 2 have the same reference numbers, and description thereof will be omitted. If atmospheric air and negative pressure are alternately supplied to the transformation chamber 108 through the atmospheric air pathway 11 and the negative pressure pathway 112 by repeated switching of the magnetic switching valve 110, the operating piston 105 is caused to rise and fall and a pumping movement can be supplied to the plunger 91. Accordingly, it is possible to adjust the amount of lubricating oil discharged by the plunger 91 by controlling the number of times the switching valve 110 is switched. The present invention is not limited to the embodiments described above, and various design modifications are possible without departing from the scope of the invention. For example, The magnetic plunger P 2 can be constructed so that an intake operation is supplied to the plunger 91 by excitation of the plunger P 2 and a discharge operation is supplied to the plunger 91 by the urging force of the return spring 97. Also, it is possible to fix the stroke of the movable core 94 and do away with the adjustment bolt 103, and if the solenoid 95 is constructed so that it is extremely resistant to weather corrosion it is also possible to do away with the cover 102. Description Of The Numerals E engine p 1 oil pump P 2 second oil pump 6 supercharging port 11 drive shaft (valve gear cam shaft) 20 pump cylinder body 24 cylinder hole 25 pump piston 26, 40 drive means (pump crankshaft, connecting rod) 83 oil pan 86 lubrication section 87 lubrication oilway for engine 89 lubrication oilway for compressor (branched oilway)

Claims: A lubrication system of a reciprocating piston type compressor for engine supercharging, comprising a pump cylinder body (20), pump pistons (25), fitted into a cylinder hole (24) of the pump cylinder body (20) at both ends of the cylinder hole (24), for defining first and second pump chambers (36 1 , 36 2 ) communicating with a supercharging port (6) of an engine (E), and drive means, arranged in an operation chamber (37) formed in a central part of the pump piston (25) for supplying reciprocating movement to the pump piston (25) using a drive shaft (11) of the engine (E), wherein, a compressor lubrication oilway (89) is branched off from an engine lubrication oilway (87) sequentially communicating with an oil pump (P 1 ), a lubrication section (86) and an oil pan (83) of the engine (E), a downstream end of this compressor lubrication oilway (89) opens out into a section, at an inner surface of the pump cylinder body (20), communicating with the operation chamber (37), and a second pump (2) having a smaller capacity than the oil pump (P1), for supplying some of the lubricating oil flowing in the engine lubrication oilway (87) to the pump cylinder body (20) side, is fitted into the compressor lubrication oilway (89). The lubrication system of a reciprocating piston type compressor for engine supercharging as disclosed in claim 1, wherein an annular groove (77) facing the operation chamber (37) is formed on an inner surface of the pump cylinder body (20), and the downstream end of this compressor lubrication oilway (89) opens out into this annular groove (77).

Inventors: TAKADA YOSHIHIRO, TSUSHIMA YUJI, WACHIGAI KAORU, TAKADA, YOSHIHIRO, TSUSHIMA, YUJI, WACHIGAI, KAORU, Takada, Yoshihiro, c/o K.K.Honda Gijutsu Kenkyusho, Tsushima, Yuji, c/o K.K .Honda Gijutsu Kenkyusho, Wachigai, Kaoru, c/o K.K. Honda Gijutsu Kenkyusho
Applicants: HONDA MOTOR CO LTD, HONDA GIKEN KOGYO KABUSHIKI KAISHA
Application year: 1999
Main classification: F02B39
IPC codes: F02B33, F01M1, F02B39, F01M11