GB2259332A - Hermetic compressor with rotary rolling piston - Google Patents
Hermetic compressor with rotary rolling piston Download PDFInfo
- Publication number
- GB2259332A GB2259332A GB9119168A GB9119168A GB2259332A GB 2259332 A GB2259332 A GB 2259332A GB 9119168 A GB9119168 A GB 9119168A GB 9119168 A GB9119168 A GB 9119168A GB 2259332 A GB2259332 A GB 2259332A
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- United Kingdom
- Prior art keywords
- crankshaft
- piston
- rolling piston
- eccentric portion
- chamber
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
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Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
- F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
- F04C18/30—Rotary-piston pumps specially adapted for elastic fluids having the characteristics covered by two or more of groups F04C18/02, F04C18/08, F04C18/22, F04C18/24, F04C18/48, or having the characteristics covered by one of these groups together with some other type of movement between co-operating members
- F04C18/34—Rotary-piston pumps specially adapted for elastic fluids having the characteristics covered by two or more of groups F04C18/02, F04C18/08, F04C18/22, F04C18/24, F04C18/48, or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in group F04C18/08 or F04C18/22 and relative reciprocation between the co-operating members
- F04C18/356—Rotary-piston pumps specially adapted for elastic fluids having the characteristics covered by two or more of groups F04C18/02, F04C18/08, F04C18/22, F04C18/24, F04C18/48, or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in group F04C18/08 or F04C18/22 and relative reciprocation between the co-operating members with vanes reciprocating with respect to the outer member
- F04C18/3562—Rotary-piston pumps specially adapted for elastic fluids having the characteristics covered by two or more of groups F04C18/02, F04C18/08, F04C18/22, F04C18/24, F04C18/48, or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in group F04C18/08 or F04C18/22 and relative reciprocation between the co-operating members with vanes reciprocating with respect to the outer member the inner and outer member being in contact along one line or continuous surfaces substantially parallel to the axis of rotation
- F04C18/3564—Rotary-piston pumps specially adapted for elastic fluids having the characteristics covered by two or more of groups F04C18/02, F04C18/08, F04C18/22, F04C18/24, F04C18/48, or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in group F04C18/08 or F04C18/22 and relative reciprocation between the co-operating members with vanes reciprocating with respect to the outer member the inner and outer member being in contact along one line or continuous surfaces substantially parallel to the axis of rotation the surfaces of the inner and outer member, forming the working space, being surfaces of revolution
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
- F04C29/00—Component parts, details or accessories of pumps or pumping installations, not provided for in groups F04C18/00 - F04C28/00
- F04C29/0042—Driving elements, brakes, couplings, transmissions specially adapted for pumps
- F04C29/005—Means for transmitting movement from the prime mover to driven parts of the pump, e.g. clutches, couplings, transmissions
- F04C29/0057—Means for transmitting movement from the prime mover to driven parts of the pump, e.g. clutches, couplings, transmissions for eccentric movement
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
- F04C29/00—Component parts, details or accessories of pumps or pumping installations, not provided for in groups F04C18/00 - F04C28/00
- F04C29/02—Lubrication; Lubricant separation
- F04C29/028—Means for improving or restricting lubricant flow
Landscapes
- Engineering & Computer Science (AREA)
- Mechanical Engineering (AREA)
- General Engineering & Computer Science (AREA)
- Applications Or Details Of Rotary Compressors (AREA)
- Compressor (AREA)
Abstract
The invention relates to a hermetic compressor with a rolling piston 50' of annular section, enclosing a crankshaft 60' having an eccentric portion 60a'. The piston 50' rolls around the eccentric portion 60a' of the crankshaft 60', all held within a cylindrical chamber of a cylinder block. The piston 50' acts in combination with a sliding vane to separate the cylindrical chamber into suction and discharge regions. These regions have lower internal pressures than the oil elsewhere inside the shell of the compressor, during most of the cycle of rotation of the piston 50'. Consequently, oil passes through small gaps between the annular end-faces of the piston 50' and the end walls of the cylindrical chamber, to lubricate, and seal, components of the compressor. By providing a rolling piston with a high ratio of external to internal diameter (greater than 1.63 but preferably less than 2.22), the radial path of the lubricating oil is increased. In this way the flow of lubricating oil is reduced, without the disadvantage of the conventional method of reducing oil flow, in which the annular faces are moved closer to the end walls of the chamber resulting in increased friction. Furthermore, the crankshaft has an end portion 60b' which has a radius R which satisfies the condition 2Rr > Rm + Ec + Re where Rr is the internal radius of the piston 50', Ec is the eccentricity of the eccentric portion @60a' of the crankshaft 60' and Re is the radius of the eccentric portion. <IMAGE>
Description
-% 7 C? 7 1 HERMETIC COMPRESSOR WITH ROTARY ROLLING PISTON The present
invention relates to a rotarH hermetic compressor of the rolling piston tHpe and of high internal pressure in the shell, low back pressure and small displacement volume, which is normallj used in small refrigerating machines.
BH rotaru rolling piston hermetic compressor with high internal pressure, it should be understood that one whose shell is submitted to the condensing pressure of the sHstem in which it is used (the so called 'high side compressor).
In rotarj rolling piston hermetic low back Pressure and high internal pressure in the shell, it occurs the Phenomenon of the passage (penetration) of the lubricating oil to the interior of the cilinder, into the suction and the discharge chambers.
The oil contained at the bottom of submitted to the high gas pressure i will be elevated bj an oil till it reaches the crankshaft is then radiaLLH displaced through the shell and nside the shell, pump or another device, from which the oil the gaps between the annular end faces of the rolling piston and the bearing covers, entering the cilinder internal chambers.
This penetration of oil at a high temperature into the cHlinder causes an the functioning and performance of the compressor the effects discussed hereinafter.
On penetrating into a high temperature, the suction chamber, the oil, at warms up the incoming suction gas] causing the increase of its specific volume and, therefore, reducing the suction chamber filling up the cHlinder capaci t1.
Thus, the gas mass which fills 2 suction chamber is redured bH the effect o '---t h e increase of the gas specific volume. Besides.this inconvenience, it should be noted that the oil volume itself which penetrates in the suction chamber takes gas filling space; however, this effect is of a quite secondari importance in relation to the heating effect.
The above mentioned problem causes a decrease in the compressor pumping capacitH as a result of the lubricating oil penetration into the cHlinder.
In turn, on penetrating into the cHlinder compression chamber, the oil will be, during a great part of the compression period, at a higher temperature than the gas temperature under compression, also causing the heating of such gas and increasing its specific volume. This phenomenon results in an increase in the work required for compressing the gas and, consequentlH, an increase in the compressor energetic consumption. This fact can be verified in figure 2 of the accompantjing drawings, where angle of rotation diagram is demonstrating that the compression faster when the oil leakage to the compression chamber increases.
considerablj drop the compressor energetic efficienci.
On the other hand, the presence of oil flow carries out two favorable functions which are fundamental to the compressor functioning.
The first one, and the most obvious one, lubrication of the movable parts involved.
The second one is the sealing of all clearances between the movable parts, thus avoiding the gas direct leakage from the interior the interior of the shell, which 1 p r e pressure raises interior of the a pressure X sented for contribute to volumetric and the lubricating i s the of the cilinder to eakage, in case of 3 happening, can be even more compressor$ in terms of capacit overheating bj the oil.
This propertij of the oil of sealing gaps between the movable parts, acts on the cilinder internal leakages (from the compressor chamber to the suction chamber at low pressure), and on the leakages compression chamber to the interior sid s h e 11.
In the more specific case of the oil whic radiallH into the ctdlinder through the ro end faces, the lubricating oil prevents leaking from the compression chambe crankshaft internal parts and from the La interior of the shell.
Therefore, the amount of oil which gets into the c1linder must be controlled at an optimum level, i.p-'' at a minimum level in order to make possible to have the sealing of the gas leakages and, at the same time, onlH a m of the cilinder. A well-known wa!3 penetrates into the rolling piston end to a minimum level between the rolling cover faces do not reach suc cancel the gains resultant reduction through said gaps.
Despite the Possibilitti of piston end gaps in a wa!j to reduction in the amount of prejudicial to the drop, than the gas from the e a f the h penetrates lling piston the gas from r t 0 the tter to the inimum gas heating in the interior to control the amount of oil that ctdlinder, bH the gaps of the faces, is to reduce such gaps up in which the losses bH friction a value to r-ompletelH from the oil flow reducing the rolling get an advantageous lubricating oil which Penetrates into the cHlinder, the obtained gain, in terms of energetic efficienci of the compressor, will 35 alwaHs be inferior to what would be reached bj the exclusive reduction of the oil flow, due to greater 4 friction loss as a result reduction of the said gaps.
In the development of the present invention, it was found out that an increase in the radial path of the lubricant oil flow through the opposite axial gaps of the rolling piston makes it possible to reduce, in at least 10%, the oil flow to the interior of the suction and discharge chambers, without substantiall!j increasing the friction losses between the movable 10 p a r t s.
From the equation that models the radial flow of oil through the rolling piston end faces (viscous flow between parallel discs) it can be noted that the flow is controlled bs the gap(& R) and b3 the thickness of the rolling piston wall or, relation:
of minor or greater 1 n ext.0 int.10 The behavior of such function can be observed in the graph of f i g. 3 The rolling piston dimensions found in the market compressors with small displacement volume internal diameter (int.0) relation between 1,40 and 1.55, def ining rolling pistons of thin walls. The invention aims thick wall those diameterlinternal approximate1H.
It can be notice about the value diameter = 1.63, the slope diameterlinternal diameter becoming graduallH softer reached 1.6.
more preciselH, bH the present an external diameter (ext.0) 1 to define as rolling pistons of which present the external diameter relation:> 1.63, bs the graph of fig. 3 that, till of external diameterlinternal of curve 1 n-1 external ite accentuated, after said value has i S q It should be remembered that the behavior of the curve that represents the oil flow toward the interior of the c!dlinder is proportionallw reflected in the compressor performance alreadi explained, i.e., the higher the function diameterlinternal diameter), the 9 n-1 (external reater will be the oil flow and worse the volumetric and energetic efficiencii oC,-,the compressor. Therefore, it is important to notice that the P('n.etration 'of i inward the cHlinder can have a reduction of at least 10% with a dimensioning of the rolling piston diameters in a waw to get a relation external diameterlinternal diameter) from 1.63 on, regarding the range commonlj used, the Polling piston being dimensioned in order to g e t a relation of external diameter/internal diameter 1.63, approximatel1. it is also important to mention that diameter relations of 1.63 on, up to nearlj 2.22 are Perfectli feasible in the production processes normallu used for ratarH rolling piston compressors and iet, there is no impediment or disadvantage in terms of the compressor performance when using such relations.
The onlj disadvantage that one could quote would be the increase of friction losses between the piston faces and the cHlindrical chamber end walls due to the increase of the contact surface. However, the increase of the friction losses does not effectivetH occur because, with the increase of the contact surface, there is a tendencu to have a reduction of the angular velocit!j of the piston over its own shaft which compensates the loss. Besides, friction is of the order of magnitude time smaller than the losses caused oil, when the gap R is the usual such loss could be neglecte It was also observed that aniwaH. certain prior such loss bi at least one the heated Therefore, art rotani 6 hermetic compressors with high displacement (higher than about in the shell (low diameterlinternal diameter relations for the rolling piston within the range of 1.63 to 2.22. However, the existence of ratarH hermetic compressors with high displacenent volume andlor low internal pressure in the shell having such dimensional rela-tion is merelH casual. There is not antd technical literature suggesting the use of said dimensional relation to obtain a reduction of the oil flow to the interior of the suction and discharge chambers.
According to the available technical literature, it can be affirmed that the hermetic compressors pr relation is a-simple cons displacement volume, whi to a preset capaciti, i thus making high the rolling piston radii, determined bj its mini strength of the materia words, it can be said with high displacement small enough to allow the eccentric and, constquentlH, an also small internal radius for the rolling piston. Thus, the external diameterlinternal diameter relation of the rolling piston can be situated within the above mentioned range onlH casuallH.
Although there are rotarH hermetic compressors with a rolling piston presenting a thick walt,it should be noted that such compressors are of the "Low-side" t1pe (Low internal pressure Nevertheless, regarding the volume 7cc) andlor low internal pres sure side compressor) present external fact of existing rotarH esenting said dimensioning equence of the fact that the ch was designed to correspond 5 high (higher than 7.1 cc), values of the citinder and whereas the mum possibl 1 which i that in hermetic compressors volume, the shaft radius is a rel-a-tivelH small radius for shaf t v a 1 U e ' u s e d.
radius is due to the In other in the shell). there are fundamental differences finalitti and functioning of a thick where:
rolling piston in pressure in the she "Low-side' hermetic compressor Claw internal pressure in the shell). In the low-side compressor, the low internal pressure in the shell does not actuate on the oil which, therefore,it is not allowed to reach the interior of the cwlinder through the gaps between the movable parts, as it occurs in the high-side in the low-side compressor, the as a sealant against the gas s, the compressed gas in the compression chamber tending to leak through the gaps between the movable parts, more specificallH between the rolling piston end faces and the bearing covers.
The flow in said gaps is, therefore, of gas leakage in the low-side compressor, and not of oil penetration as it occurs in the high-side 7 compressor. Thus, oil does not act leakages through the gap compressors.
Reducing the gaps 0 thickness in the or increasing low-side the rolling piston compressors has the finalit3 of avoiding the gas leakage in the compression chamber, and not of controlling the problem of oil flow to the interior of the citinder, as it occurs in the high-side compressors. Thus, the the rolling completelli finaliti of increasing the thickness of piston in both tHpes of compressors is different. As the internal diameter of approximateltd the same as eccentric portion of the relation can be represented a 1.63 RR o r RR R r Re RR = ext-ernal radius of the rolling piston the rolling piston is the diameter of the rankshaft, the desired follows (see fig.4b):
2.22 ( 1) is 8 Rr = internal radius of the rolling piston Re = radius of the crankshaft eccentric portion As the external radius RR of the rolling piston is determined in relation to the culinder diameter that is designed for the compressor, the relation (1) above can be achieved bts changing the values of the piston internal radius Rr and, consequentlw, the radius Re of the crankshaft eccentric portion.
In the known rotarid hermetic compressors (having a displacement volume above 7 cc), presenting the dimen's.ional relation (1) above, the internal radius R r o f the r a 11 Phs piston o r r a'd i u s Re o f the crankshaft eccentric portion) is generalts enough to allow the following dimensional relation:
Rr z Re = Ec + Rs (2) large where:
Ec = eccentriciti of the eccentric portion Rs = radius of the compressor shaft The dimensional relation (2) above is shown in figure 4a, though this prior art solution does not necessarilH present the dimensional relation (1) simultaneouslH.
When the compressor presents the dimensional relation (2) above, the radius Rm of the shaft end portion can be maintained equal to the radius Rs of the crankshaft, i.e., Rm = Rs, without causing anw problem of assembling the rolling piston on the eccentric portion of the crankshaft, as illustrated in fig.4a, where the contour of the eccentric portion is tangent to the crankshaft remainder contour.
Nevertheless, in the rotarH hermetic compressors with small displacement volume (less than 7cc) and high internal pressure in the shell, the reduction of the internal radius Rr of the piston (or radius Re of the eccentric portion), in order to achieve a radial extension of the piston wall within the relation (1), 9 can make it impossible to have, due to the eccentricitH Ec required in the compressor design and to the minimum diameter required for the shaft, both dimensional relations (1) and (2) simultaneouslH. In 5 these prior art compressors, the dimensional relation (1) can onlj be obtained in conjunction with the following dimensional relation (fig.4b);
Hr ( Rm + Ec + Re (3) In this situation, the contour of the crankshaft eccentric portion is not tangent to the crankshaft contour anumare, becoming secant to the latter, avoiding that the crankshaft eccentri rolling piston be mounted at the c Portion.
Thus, it is an object of the present invention to provide a rotaru rolling piston' hermetic compressor with high internal pressure in the shell and small displacement volume (Less than about 7cc), which presents a lubricant oil flow to the interior of the 20 culinder, considerablu reduced in relation to the known solutions, without causing anj substantial between the compressor movable lj between the rolling piston covers and between the portion of the driving increase of friction parts, more specifical end faces and the bearing piston and the eccentric crankshaft.
The hermetic compressor with rotarjrolling piston, of small displacement volume and low back pressure comprises: a hermetic shell submitted to high pressure; a cwlinder shell and has crankshaft block that is housed inside the an internal cilindrical chamber; a having an eccentric portion which is close to an end portion of the crankshaft; a rolling piston which is assembled around the eccentric portion of 35 the crankshaft, in order to rotate inside the cilindrical chamber; end walls that close the.
opposite axial ends of the cHlindrical chamber, said chamber being internallH divided bs the rolling piston in a suction chamber, whose internal pressure is substantiallH lower than the internal pressure of the shell, and in a compression chamber presenting an internal pressure substantiallti lower than the internal pressure of the shell during most of the compression CHcle; and an axial gap for the passage of lubricant oil between said end walls of the c1lindricaL chamber and the opposite end faces of the rolling piston.
According to the present invention,the rolling piston is built inorder following dimensional relations:
1.63 RE 2.22 1 R r Rm + Ec + Re (4) whe re itH of the eccentric portion to simultaneouslb present the 2Rr Ec = eccentric Re = radius of the eccentric portion RR = external radius of the rolling piston Rr = internal radius of the rolling piston Rm = radius of the crankshaft end portion Rs = radius of the crankshaft so as to increase the radial path of the lubricant oil through said axial gaps. In the cases where the diameter of together with the dimensional relations above, allow the relation:
Rr -- Re = Rs + Ec (2) the shaft, ( 1) and (4) the assemblu of the rolling piston to the eccentric portion of the crankshaft can be made bj maintaining Rm = Rs, with the contour of the eccentric portion being kept external and tangent at a point in relation to the contour of the crankshaft.
In the cases where the dimensional relations (1) lead to the relation:
R r 2: Re ( Rs the mounting of 11 + Ec (5) the rolling piston to the eccentric portion of the crankshaft can onlH be made btd reducing the diameter of the crankshaft end portion so as to have Rm Rs and make possible the dimensional relation (4). In this situation, the contour of the crankshaft eccentric portion is secant in relation to the crankshaft.
In another embodiment of the invention, the crankshaft end portion is not provided with a bearing. In this case, the axial end wall of the ctilindrical chamber faces the crankshaft bodj and is defined bj the plate of a respective bearing that is attached to the cHlinder block, whereas the opposite axial end wall is defined bH a plate which is attached to the adjacent face of the c3linder block. In this constructive solution, the rolling piston is designed in such a wai to present the dimensional relation (1).
The invention will hereinafter be described, reference to the appended drawings, in which; Fig.1A shows a partial longitudinal section view of a rolling piston rotarH compressor of the tHpe used in the present invention; Fig.1B shows a cros to line B-B of fig.l; Fig.2 shows a diagram of the compression produced in terms rolling piston; Fi.9.3 illustrates a representative function of oil radial flow through piston faces X the thickness of the of the rotation an pressure 9 1 e of the graph of the t h e ro 11 ing piston annular wa 11 ' Fi.9.4A illustrates a side view of the crankshaft rolling piston set of the prior art;
12 Fig.4B illustrates a side rolling piston set built according to a embodiment of the invention, in which the dimens (1) is obtained bH reducing the internal rolling piston and the diameter of the relation radius of the whole shaft; Fig.4C adapted which a 10 portion diamet r a is view of the crankshaft f i r s t ional shows the set of the previous figure as to a second embodiment of the invention, in diametral reduction is made at the end of the shaft, in order to achieve the 1 relation (1); According to figures 1A and IB, the compressor of the 20 present invention includes a shell 1 fastening suction 2 and discharge 3 tubes and housing a cHlinder block 4, in whose interior a cHlindrical chamber is defined which houses a rolling piston 5 that is mounted on a crankshaft 6 driven bH an electric motor composed bH a stator 7 and a rotor 8. This compressor is of the tHpe which presents high internal pressure in the shell, low back pressure and small displacement volume. Inside the shell, an inlet end of the discharge tube 3 is opened.
crankshaft 6 is supported on a a sub-bearing 20, each flange 10a and 20a fixed faces of the cHlinder blo cilindrical chamber walls i rolling piston 5 is displaced In the illustrated exampl The and e 1 a main bearing 10 one embodHing a plate or against one of the axial end ck 4, in a wai to define the n which interior the discharge muffleC 1 13 chamber 13 is provided next to the sub bearing 20 external face, so as to receive the compressed gas inside the cHlinder, the sub bearing plate 20a (or the ctilinder block wall 4 in case of absence of such bearing) being provided with an orifice 22, with its outlet end defining an annular valve seat against which is seated a known reed valve 30 internal to the discharge muffler chamber 13. Still according to the basic constructio in figures 1A an 1B, the main bear provided with a radial channel 11 w being immersed in the lubricating the bottom of the shell 1 and wit being opened to an oil pump or device, defined around th communication wit 14 provided along conduct the lubri axial gaps of the As illustrated i presents windows the shell and for slot where a together with th rolling piston 5 a suction chambe one being fed citinder block the internal en illustrated ng plate 10a is th its lower end oil OL stored at h its upper end another pumping e crankshaft 6 and in fluid h longitudinal superficial grooves the crankshaft 6, in order to ating oil to the bearing and to the rolling piston 5. figure 1B, the cjlinder blo 4a for internal pressure balance in Lubricating oil passage and also, a liding vane 9 is incased which, external cHlindrical face of the divides the culindrical chamber in and a discharge chamber, the first through a channel 4b made on the and maintained in communication with of the suction tube 2.
Figure 4A illustrates a conventional construction of a crankshaft 6 including an eccentric portion 6a for driving the rolling piston 5 and an end portion 6b for journalling inside the sub bearing 20; usualli, this specific crankshaft portion has the same diameter of the rest of the crankshaft bodH. In this solution of the prior art, corresponcf'ing to the c k 4 14 arrangement illustrated in figure 1B, the eccentric piston 5 presents an annular wall thickness accor ' ding to the relation RR 1 Rr ( 1.60, making the oil path, represented bi the arrows in figure 1B, be short enough to allow a prejudicial amount of lubricant oil to penetrate inside the c!dlinder.
Figure 4B illustrates a crankshaft 60 with its portion 60a driving a rolling piston 50 with annular wall thickness according to a relation RR 1 Rr) 1.63. This dimensional relation was reached bH reducing the diameter of the entire crankshaft 60, jointlu with a corresponding reduction of the internal diameter of the rolling piston SO and, consequentlti, of the eccentric portion 60a. In this particular case, it was possible to reduce the diameter of the crankshaft 60 (including its end portion 60b), in order to maintain the dimensional relations:
RR R r 2 Rr Figure 4C illustrates eccentric portio having an annular relation RR 1 Rr 1.63 > Rm + Ec + Re (4) Rs = Rm crankshaft 60', with its 6Oe' driving a rolling piston SO' wall thickness according to the 1.63. This dimensional relation, which is also within the limits defined in the present invention--,-- was reached bs reducing the internal diameter of the rolling piston SO' and, consequentlu, the diameter of the eccentric portion 60a'. In this case, the reduction of the shaft diameter is not possible,for example,due to design reasons,and the reduction of the eccentric portion 60a' maH lead the set to the following dimensional relations:
If the Rm Rs a s relation w is Re ( Rs + Ec value is kept in fig. 4B), ill occur:
2 Rr ( Rm + Ec + Re condition in which piston 50 ec is (5) the same as the Rs value (Rm the following undesirable (3) indicating a the internal diameter of the rolling is smaller than the joint contour of the centric portion 6Oe' and the end portion 6W of the crankshaft 60' ( in this case, the shaft end portion is kept with the same diameter as the crankshaft; see dashed lines in figure 4C.).
Thus, as illustrated in figure 4C, the diameter of the end portion 6W of the crankshaft 60' is reduced so as to achieve the following dimensional relation:
RR Rr 2 Rr where Rm ( Rs.
4D ill k 1.63 ( 1) Rm + Ec + Re (4) Figure ustrates for the crankshaft large thickness of 1 presents an eccentric reduced diameter, the end portion havi completelH eliminated and, consequentlj, bearing 20 as well. In this case, the axial end well of the cilindrical chamber defined bH a simple closing plate which is to the adjacent face of the cHlindrical block.
another possible 600 which, as a he rolling portion construction result of the piston 500 wall, 600a with a verH ng also been the sub corresponding can be fastened - Hermetic internallH suction substantial s h e 11, and internal internal compression
Claims (5)
1.63 _Ell 2. 22 1 Rr 2Rr Rm + Ec + Re (4) where: Ec = eccentricitti of the eccentric portion Re = radius of the eccentric portion RR = external radius of the rolling piston Rr = internal radius of the rolling piston Rm = radius of the crankshaft end portion Rs = radius of the crankshaft in order to increase the radial path of the Lubricant oil through said axial gaps.
2 - Hermetic compressor, according to claim 11 wherein the rolling piston (50) and crankshaft (60) set simultaneouslu presents the following further dimensional relations:
Rr z Re = Rs + Ec (2) Rm = Rs the contour of the eccentric portion (60a) of the crankshaft (60) being external and tangent to the contour of the crankshaft (60).
3 - Hermetic compressor, according to claim 1, wherein the rolling piston (50') and crankshaft (60') set simuLtaneouslH presents the following further dimensional relations:
Rr R e R s + E c Rm Rs the contour of the eccentric portion (60a') of the crankshaft (60') being secant to the contour of the crankshaft (60').
4 Hermetic compressor, according to claim 1, wherein the axial end walLs of the cilindricaL chamber are each defined bH the plate (10a,20a) of a respective bearing (10,20) which is fastened to the cHLinder black (4).
5. A hermetic compressor comprising a rotary rolling piston of annular section enclosing and rotatable around an eccentric portion of a crankshaft; the piston and eccentric portion being held within a cylindrical chamber of a cylinder block; the piston and an externally abutting sliding vane cooperating to separate the cylindrical chamber into suction and discharge regions; the cylindrical chamber having at least one end separated from a neighbouring annular end-face of the piston by a gap through which in operation lubricating oil flows, the piston having a ratio of its external to its internal diameter of at least 1.63, the crankshaft having an end portion extending into the gap, the end portion having a radius R. which satisfies the condition R. + E. + Ra < 2R, where R, is the internal diameter of the piston, E. is the eccentricity of the eccentric portion of the crankshaft, and R. is the radius of the eccentric portion.
c X 1
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| US07/717,691 US5226797A (en) | 1989-06-30 | 1991-06-19 | Rolling piston compressor with defined dimension ratios for the rolling piston |
Publications (3)
| Publication Number | Publication Date |
|---|---|
| GB9119168D0 GB9119168D0 (en) | 1991-10-23 |
| GB2259332A true GB2259332A (en) | 1993-03-10 |
| GB2259332B GB2259332B (en) | 1994-12-14 |
Family
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Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| GB9119168A Expired - Fee Related GB2259332B (en) | 1991-06-19 | 1991-09-06 | Hermetic compressor with rotary rolling piston |
Country Status (2)
| Country | Link |
|---|---|
| DE (1) | DE4137363A1 (en) |
| GB (1) | GB2259332B (en) |
Families Citing this family (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| DE602004027781D1 (en) | 2003-09-30 | 2010-08-05 | Sanyo Electric Co | rotary compressors |
Citations (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| GB951837A (en) * | 1962-01-10 | 1964-03-11 | Pablo August | A machine capable of operating as a compressor or pump |
| GB1560438A (en) * | 1976-12-13 | 1980-02-06 | Borg Warner | Cooling system for rotary hermitic refrigerant compressor |
| GB2222205A (en) * | 1988-06-08 | 1990-02-28 | Egan Michael J | Orbiting eccentric pumping device |
-
1991
- 1991-09-06 GB GB9119168A patent/GB2259332B/en not_active Expired - Fee Related
- 1991-11-13 DE DE19914137363 patent/DE4137363A1/en not_active Withdrawn
Patent Citations (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| GB951837A (en) * | 1962-01-10 | 1964-03-11 | Pablo August | A machine capable of operating as a compressor or pump |
| GB1560438A (en) * | 1976-12-13 | 1980-02-06 | Borg Warner | Cooling system for rotary hermitic refrigerant compressor |
| GB2222205A (en) * | 1988-06-08 | 1990-02-28 | Egan Michael J | Orbiting eccentric pumping device |
Also Published As
| Publication number | Publication date |
|---|---|
| GB9119168D0 (en) | 1991-10-23 |
| GB2259332B (en) | 1994-12-14 |
| DE4137363A1 (en) | 1992-12-24 |
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Legal Events
| Date | Code | Title | Description |
|---|---|---|---|
| PCNP | Patent ceased through non-payment of renewal fee |
Effective date: 20030906 |