DE2434873A1 - ROTARY SCREW COMPRESSOR WITH OPTIMALLY LOCATED POINT FOR INJECTING LIQUID COOLANT - Google Patents

Description

25 536 t / wa

DUNHAM-BUSH, INC., WEST HARTFORD
CONN., / USA

Rotary screw compressor with optimal
arranged point for injection of a liquid
Coolant

The invention relates to a rotary screw compressor
for a cooling system and in particular such a compressor in which a liquid coolant enters the working space
of the compressor is injected to cool the gaseous refrigerant working medium during its compression.

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The invention relates to the field of positive displacement Rotary screw compressor of oil injection or dry construction, where the injection point for the liquid coolant is optimally arranged. Screw compressors of the type in question, for compressing an elastic working medium, generally have a housing a working space which is formed by two parallel and penetrating bores. At one end of the There is a low-pressure or suction opening in the working space and a high-pressure or outlet opening at the other end.

In the two bores, two intermeshing screw rotors are rotatably arranged and furthermore, if desirable to form a slidable relief valve part of the wall of the working space. The axial position of the The relief valve determines the capacity of the machine.

To avoid the use of timing gears to synchronize the rotors, a liquid such as for example oil, is injected into the compressor. This oil serves to seal the free space between the rotors and between the rotors and the housing, thereby reducing gas backflow, and at the same time to cool the working medium to be compressed.

For rotary screw compressors that are part of a cooling or form air conditioning systems and in which the working medium is a typical coolant, has already gone on proposed to inject the coolant in liquid form into the working chamber of the compressor for the purpose of cooling.

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For this purpose, the liquid coolant which is under high pressure is branched off from the condenser in the cooling circuit, this being the case is essentially at compressor outlet pressure. A controlled injection of a liquid coolant in the mentioned, th way is described in the German patent application P 23 43 33.

The cooling of the compressor by means of injection of the cooling liquid means that the usual oil cooler can be dispensed with can. The dissipation of the heat from the oil cooler usually represents a not insignificant percentage of the total heat dissipation due to the amount of circulating oil. With the So coolant injection can be an oil cooler as well as the bypass associated management and control means.

The definition of the place within the compression cycle in which the liquid coolant is fed to the working chamber of the compressor is therefore extremely difficult as the load on the compressor changes. In particular ^ which poses a problem in cases where the screw compressor does not work with a constant compression ratio. The screw compressor forms a component of the cooling or air conditioning system and if the loads on the system change, the operating compression ratio changes with a change in suction pressure, which also changes the compressor discharge pressure.

The aim of the invention is therefore to provide a rotary screw compressor to improve the genre mentioned in the preamble of claim 1 in such a way that the point at which the liquid coolant is injected into the working chamber,

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is optimally adapted to the changing system operating parameters.

According to the invention, this is achieved in that the injection opening for the liquid coolant is arranged at a point relative to a specific working chamber *, which by the current pressure relationship between the diverted liquid coolant and the pressure of the compressor working medium is determined during the compression in the chamber, so that the cooling of the working medium during the compression by continuously injecting the diverted liquid coolant at the point regardless of over- or under-compression of the working medium and without significant impairment of the compressor performance is guaranteed.

The procedure is used to determine the optimal injection site proceeded so that you get the minimum built-in volume ratio of the compressor determines, determines within which borehole the injection should take place, 'the minimum and determines the maximum operating compression ratio of the compressor, using the determined parameters. the pressure-envelope angle diagram for the rotor of the selected bore in relation and a suitable location from the diagram for the injection port with respect to the angle of wrap from the suction inlet of the screw rotor for the selected bore determined, whereby a continuous injection of the liquid coolant into the work space independent of a Under- or over-compression of the working medium within the results in limits set by the determined parameters without significantly affecting the compressor performance will.

of the compressor

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An exemplary embodiment of the invention is explained in more detail below with reference to the drawing. Show it:

Fig. 1 is a sectional, partially schematic view of a rotary screw compressor with the coolant injection system according to the invention;

2 is a schematic, perspective view of the meshing positive and negative screw rotors the compressor of Figure 1;

3 shows a diagram of the pressure curve in the screw compressor according to Fig. 1 and a two-dimensional representation of the change in volume of the compressed Working medium from the suction to the outlet side;

FIG. 4a shows a PV diagram for the compressor according to FIG. 1 under ideal conditions; FIG.

FIG. 4b shows a PV diagram for the compressor according to FIG. 1 with over-compression; FIG.

FIG. 4c shows a PV diagram for the compressor according to FIG. 1 with under-compression; FIG.

5 shows an inlet-side end view of the intermeshing screw rotors with illustration the formation of a closed pocket;

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6 shows an outlet-side end view of the compressor according to Fig. 1 with representation of the angular relationship between the screw rotors and one of the chambers, which is sealed off from the suction and discharge sides of the compressor;

7 shows a pressure-envelope angle diagram for the positive rotor in the end plane on the outlet side the compressor of Figure 1;

8 shows a pressure-envelope angle diagram for the negative rotor in the end plane on the outlet side of the compressor according to FIG. 1

To fully understand the parameters associated with the Determining the optimal injection site for a liquid refrigerant in a screw compressor is also considered a visual understanding of the way in which the volume of the between the helixes increases is necessary and gas chambers formed in the housing changes. The rotary screw compressor represents a positive displacement compressor, similar to the reciprocating compressors in which a piston moves back and forth in a cylinder and the gas pocket has a maximum size at the end of the suction stroke. In contrast, the chamber in such piston compressors has its smallest volume when the piston is on the compression stroke has reached its top dead center. Such a change in volume can easily be imagined. at In typical rotary sliding vane compressors, the rotor is also arranged eccentrically in a cylindrical housing in which he turns. The sliding wings can be moved radially in the rotor,

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wherein between the housing and the rotor and a subject fenden wing pairing a chamber is defined. The chamber is large on the suction side of the machine and decreases considerably after turning it by about 180 ° towards the outlet opening. The change in volume of such a chamber, i.e. its reduction in size from the suction to the outlet side, can be visualized also easily two-dimensional by a simple cut at a right angle to the axis of rotation of the rotor easily illustrate.

Although screw compressors have fewer moving parts and are based on a fairly simple principle, leaves The change in volume of the individual gas chambers is no longer shown in a simple two-dimensional representation illustrate that the gas chambers are determined by the helical screws and the housing. Rather is here too consider the third dimension ^.

Fig. 1 shows in two-dimensional and partially schematic Shape a typical screw compressor in cross section. In figurative terms, the compression of that which is in the form of a gas moves Working medium from the low-pressure suction or inlet side of the compressor to its high-pressure or outlet ^ · page. For this purpose, the compressor, generally designated 10, comprises a three-part housing with parts 12, 14 and 16, the housing part 12 forming the suction side of the machine and being provided with an inlet passage 18. On the housing part 12 adjoins the middle housing part 14 in which the two intermeshing helical screw rotors are included; only one rotor is shown in block form at 20 in the figure. The housing part 14 therefore forms

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the working space in the form of two penetrating bores, the bore 22 containing the rotor 20 and 'the Suction passage 18 into a low-pressure suction port at the end The suction volume is shown schematically in a rectangular shape at 26 by the lines 28, 29, 31 and 32 indicated. A combined axial and radial high pressure outlet The opening 3 4 at the opposite ends of the rotors opens into an outlet passage 36 which is the main Area of the housing part 16 occupies. Each rotor, such as rotor 20, is rotatable by bearings 38 and 40 held. The bearings 38 and 40 take the protruding ends of the respective shafts 42 in the housing parts 12 and 16 respectively. In the present case, the rotor 20 compulsorily driven by a drive shaft 44 which can form an integral part of the rotor and with coupled to a prime mover, not shown. For the invention pertinent, although with rotary screw compressors Common is a relief valve, generally designated 46, which is reciprocable in a bore 48 of the middle housing part 14 is arranged. The valve 46 forms part of the working space, which is otherwise through the bores, such as the housing bores, and the rotors themselves, is defined. In the present case, the relief valve 46 is crenellated at its outlet end or recessed leading edge 4 9 defining the radial outlet portion of the outlet opening 34 of the compressor forms. The relief valve 46 has such an axial length and operates with a stationary rather than a fixed one Stop for the valve 46 acting wall area 50 of the housing part 14 together in such a way that a bypass or Relief opening 52 is created, which is directly to the suction

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or inlet passage 18 leads. A shaft 54 connects the valve 46 to a piston 56 which is reciprocable is arranged in a cylinder 58. The cylinder, 58 is on Housing part 12 attached, the piston 56 between the fully extended and the dashed position by the controlled, not shown supply of a print medium, '. such as one from the compressor's lubrication system diverted lubricating oil, can be moved in order to move the relief valve between the load and relief position according to the load requirement of the cooling system in this way.

Fig. 1 shows in two-dimensional form the schematic Change in volume of the gas chambers formed between the coils of the rotors and the housing. Best of all this can be seen perhaps from Fig. 2, which shows the considerable change in gas volume from the left-hand upper end of the Rotors pointing to the outlet at the lower right end of the rotors. Both the positive rotor 20 (male rotor) and the negative rotor 21 (female rotor) shown in dashed lines as a penetrating cylinder, with the penetrating cylinders Cylinders are characteristic of the interlocking projections and indentations on the rotors in question. In the schematic perspective illustration according to FIG. 2, the relief valve 46 is in that shown in FIG. 1 Full load position shown. Immediately to the right of the ■ relief valve 46 is the volume of the compressed gas, which moves axially as well as radially to the outlet passage opens, indicated at 60. It takes on an irregular shape due to the configuration of the rotors at the outlet end.

In the two-dimensional image of FIG. 1, the form

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Rotors meshing with one another have individual gas chambers or pockets which are separated from one another by vertical sealing lines, for example line 62. The chambers are designated with Roman numerals I, II, III, IV, V, VI and VII; the chambers I, II and III are open to the suction side, the • vertical sealing line 62 'represents the suction-side shut-off, ie this line intersects the lines 30 and 32 at the point 64, which corresponds to the same point 64 of FIG in which none of the right-hand gas chambers is open to the suction side. The compression takes place in the downstream direction, ie to the right of the shut-off point 64 and the sealing line 62 '. In the schematic representation, the compression takes place within the gas chambers IV, V and VI, the gas chamber VII opening towards the outlet. The sequence of suction, compression and discharge is indicated in the longitudinal direction by arrows S, C and D in both FIGS. 1 and 2. As to the relief valve 46, so this sequence remains as shown _r when the valve is not present. When the relief valve 46 is moved away from the extended full load position, where it hits the housing area 50 according to FIG. 1, a certain part of the gas normally involved in the compression process is guided back to the suction side of the machine via the bypass opening 52 and the bore 48 and arrives thus in the suction passage This reduces the amount of compressed gas, as indicated schematically in Fig. 1 by the position of the left end 66 of the relief valve 46 relative to the diagonal 30, which the change in the volume of the chambers IV, V, VI and VII represents during compression. As can be seen from the further figures, the size or the volume of the gas chambers IV to VII are reduced and thereby

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a compression of the gas contained therein is achieved when the chambers move one after the other from the suction to the outlet zone move. Instead of chamber I, a new one occurs as soon as .erstere to position II as a result of the rotation of the rotors in the manner shown in Fig. 2, which is a result of the direct meshing of the helically extending elevations of the rotors in question.

From Figs. 1 and 2 it can further be seen that the positive and negative rotors when they are inside a housing with fixed suction and outlet openings 24 and 60 are inserted, gas passages and gas chambers between their coils and form the housing, and so absorb the suction gas. This will then compressed due to the reduction in volume of the gas chambers, since the interengaging and Rotate matching rotors synchronously creating a gas pumping and compression effect. Thereby -that Coolant gas moved from the suction to the outlet side at opposite ends of the housing. Because of the complex It is difficult to define the sealing lines and the conveying movement within the compressor housing. However, this also hinders attempts to find the narrow area of the optimal position of the injection opening for the liquid To define coolant, which is the aim of the present invention.

A gas chamber created by the intermeshing positive and negative rotors and the housing part 14 is more clearly shown in Figs. The positive rotor rotates at a speed 1 1/2 times that of the negative rotor. A horizontal one

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The center line extends through the axes of both rotors, see FIG. 5, and intersects the contact point between the projection 1 of the positive rotor 20 and a point between the projections 1 and 2 of the negative rotor provided with six projections 21, when viewed from the suction or inlet end. The gas chambers thus formed surround the negative rotor 21 in the direction of the outlet side of the machine at an angle O ^ 1 of 200 °, and in a corresponding manner the chamber area of the positive rotor 20 assumes an angle φ 2 of 300 ° around the rotor axes, as indicated in the illustration according to FIG. 6 relating to the outlet end.

The screw compressor shown in the figures has a fixed, built-in volume ratio of V. = V _c

where V _{c is} the volume on the suction side and V _{n is} the volume

ü L)

on the outlet side. The pressure ratio P. of the compressor is in this case equal to V., where k is a factor for the specific heat of the medium to be compressed. The working medium for the compressor can be a coolant such as R22 with k = 1), 15 or ammonia (NH ₃ ) with k = 1.29.

As can be seen more clearly from Fig. 4a, 4b and 4c, in which

P. = —ρ—, the system compression ratio xs C / R ¹ S is correct

coincides with the compression pressure ratio under ideal conditions, i.e.

^P D
P. = -5- = G / R, see Figure 4a.

^x ■ S

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In Fig. 4b, a situation is shown in which there is over-compression and the compressor pressure ratio P ^ is greater than the system compression ratio C / R = P / P. The over-compression and the resulting loss of work is indicated graphically by the hatched area in the PV diagram. Opening a certain chamber in the compressor towards the outlet end causes the pressure within the chamber to drop to the much lower outlet pressure.

In Fig. 4c the effects of under-compression are shown, in the

P.
P. becomes smaller than - = - = .C / R. The hatched area

¹ s

indicates the loss of work due to under-compression, whereby the outlet has to further compress the working medium in the compressor at a point at which the chamber opens towards _{the outlet pressure P 0.}

The effect of under-compaction and over-compaction on the injection site is best illustrated by this from Fig. 3 recognize in the in-a common diagram the "pressure along the compressor from suction to discharge and the change in volume from the suction to the outlet side of the compressor are shown schematically.

Fig. 3 further shows the over- and under-compression by the dashed lines P ^ and P ^ ,.

Over-compression occurs when the suction pressure increases from P * to P _q2 . Then the internal due to the

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built-in P. reached outlet pressure in chamber VI P ", when it opens to the outlet port 34, as the built-in pressure

^P D2

P ₁ - = ~ ü is. Hence, Δ = P - ρ, which is a

F _s2 P OZ Di

Expansion of the gas to the discharge end and a drop caused, which is indicated by the drop in curve P ₂ to the distributor pressure P.

Under-compression occurs when the suction pressure of P ₁

falls to P-, where then due to the fixed P: = - ^

^{SJ x P} S3

the outlet pressure present at the outlet opening '34 is equal to P _n -. will. As the manifold pressure is higher; becomes A _p = P ^ ₁ -P _n -. and thus further compressed by the return flow. The effect of over-compression and under-compression and the restriction created by the relief valve on the optimal liquid injection zone can be seen from FIG. 3. The injection zone or the injection window for the liquid coolant is the zone between the two vertical arrows I ₁ and ~ L ~ r , the arrows of a two-dimensional schematic representation of the compression of the working medium being superimposed in the upper part of FIG Working medium in the lower part of the figure. Are reproduced. The injection zone for the liquid coolant, which is partially defined by the arrow I ₁ , lies to the right of the suction sealing line 62 for the gas chamber IV facing the front tip, approximately one sealing line away from the suction shut-off line 62.

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Although the injection zone lies within the compression area C of the compressor according to FIG At the high speed of rotation of the rotors, the kinetics of the machine when injecting the liquid into the chamber requires that the chamber be immediately sealed off from the suction side, with the gas in the chamber beginning to compress as the chamber moves to the point where it is it opens towards the outlet side of the machine. In a similar way, the vertical arrow I “indicates the other possible limit with regard to the angular position of the injection opening for the liquid coolant on the machine, ie the injection would have to stop approximately before the point at which the chamber opens to the outlet side. If the built-in ratio P. corresponds to the operating pressure ratio C / R, the theoretically maximum _{zone defined by the vertical arrows I 1} and I ₂ can apply. However, with over-compression this zone narrows because the pressure of the liquid refrigerant, defined by the compressor discharge pressure, is approximately equal to ^P DM ^ ^st 'where P ™ represents the discharge manifold pressure (the discharge manifold pressure is generally equal to the refrigeration system condenser pressure from which the liquid coolant is supplied for injection). The coolant pressure must be higher than P _n -, where P _n - in the pressure chart according to FIG .. higher suction pressure P- is present. Assuming that the injection is at the high pressure end of the

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Injection zone takes place where the pressure of the working medium _{has reached the value P 1} due to the overcompression, no injection takes place if P _{01 is} greater than P '. Maintaining a continuous flow of coolant is, however, essential for adequate cooling and for the operation of the machine. The injection of the liquid coolant is, however, subject to pressure variations as a result of the internal gas pressure at the injection site, and this causes a change in the coolant flow. As can be seen from FIG. 3, there may be a drop in suction pressure, which leads to an under-compression, which in turn requires a back-flow compression to P _M.

From the above description it follows that numerous factors are decisive for the determination of the optimal injection position for different built-in operating compression ratios for the screw compressor. As a result, various investigations must follow. First, it must be determined what the minimum built-in ratio in terms of either axial or radial compression is possible for the particular compressor. For example, if V-. raflial = 3.0 and V. axial = 3.3, the minimum V. = 3.0 _f so that this value is decisive for further calculations.

Then, in the second step, it must be determined in which rotor bore the liquid is to be injected.

Injection can be "radial" into a selected rotor bore or take place axially in an end surface.

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The next step is the maximum and minimum compression ratio, which are determined by the liquid coolant or another working medium, and others Parameters related to the pressure and the temperature of the working medium on the inlet or suction side as well as on the outlet side of the machine. For example, if R22 is used as the working medium, the minimum compression ratio is 2.7 based on a minimum of 4.4 ° C saturated suction and 4O.6 ° C saturated discharge, while for the same refrigerant at -6.8 ° C saturated suction and 63 ° C saturated discharge the compression ratio C / R assumes a maximum of 6.5.

The following is a reference to the pressure curves according to FIGS. 7 and 8 for a specific compressor and a specific working medium Referenced. Depending on whether the injection is in the bore of the positive rotor or in the bore of the negative Rotor takes place, the next step is to follow horizontal lines in the pressure curve in question Enter positions:

at C / R Max ■ = maximum compression

relationship

V. = minimal built-in

^x ratio

C / R min ^; = minimum compression ratio

With these parameters, the angle at which the injection is to take place can then be picked up. The injection should not take place at an angle at which the injection port the minimum built-in pressure ratio of the

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Compressor is exposed. Assuming the negative hole has been selected for injection and a If the minimum built-in ratio V. is 3.0, the injection should never be at an angle greater than this than 286 °.

If the injection angle exceeds 28 6 °, no injection takes place more, because at 286 °, due to the compression, there is a pressure of the working medium that is greater than the pressure of the to injecting liquid coolant is.

On the other hand, if the injection takes place between 226 ° and 286 °, the injection is intermittent and the other way round Injection flow can occur when the compressor is near its lowest, design-determined Compression ratio 2.7 according to step 3 works. This can damage the injection modulating valve and restrict the injection flow enter the liquid. Under 210, however, an injection is guaranteed because the pressure of the liquid The coolant of the line relating to the liquid coolant is far above the pressure in the special chamber, into which the injection is made.

Based on these considerations, the lower end of the over-compression range (210 ° to 226 °) forms the injection window or the zone for the injection of the liquid coolant into the bore of the negative rotor.

If you take an injection site at an injection angle of 215 °, the point is above the outermost one

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Minimum, i.e. this prevents an injection into a chamber that is open to the suction side of the machine. There further Injection at low compression ratios is not an absolute necessity, would also be one intermittent flow may be sufficient when operating at low compression ratios. If the Injection at an angle between 226 ° and 286, the injection is intermittent, i.e. it only takes place during one Part of the rotary movement takes place, while with further rotary movement the pressure in the chamber corresponds to that in the bleed line (bleed line), so that the flow of the liquid coolant in the injection line is reversed and the Net flow becomes minimal. A flow reversal in the injection line can also occur when the compressor is close at its lowest compression ratio determined by the design, i.e. at 2.7, works. It must be pointed out that as the Liquid injection to the chamber pressure, the greater the proportion of expanded coolant that will find its way back ' to the suction side and reduces the volumetric efficiency of the compressor. The closer the injection to the inlet towards the end, the greater the impairment of machine operation, so that within the injection window from 210 ° to 226 ° there must be a compromise. ·

However, if the machine is operating at a low compression ratio, for example, at the minimum ratio of 2.7, an intermittent flow is obviously sufficient at low compression ratios for sufficient cooling of the machine. In the range from 210 ° to 286 ° represents the saturation pressure, which is the condenser

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or receiver pressure is _r an acceptable pressure value above the pressure in the injection port. The measurable pressure in the injection port "corresponds annähernd'dem minimum theoretical pressure in the injection port, as by the compressor parameter plus a third of the differences between the minimum and maximum pressure 8, the available Al? represents the outlet pressure minus the pressure in the injection port. This information applies under the assumption that there are no pressure losses in the outlet line, the injection line and the TX valve.

For example: AP at the injection port with a minimum compression ratio for R22 at 4.4 ° C / 40.6 ° C would be 4.15 kp / cm ² and at -6.8 ° C / 63 ° C 18.15 kp / cm ² . The determination of Δ P is carried out as follows: The injection pressure for the liquid coolant must have an acceptable value above the pressure at the injection opening in order to ensure continuous injection. The measured opening pressure P follows the relationship (P min. + P max - P min ) where then

Δ P (for the liquid to be conveyed) (P _ P). To the Example:

P ₁ = P _{min (} , ₌ - ,, 5P ₅₊ ( ³ .S)

= 1.5 P _s + .5 P _s = 2.0 Pg

(a) 4.4 ° C / 4O, 6 ° C Pp-P ₁ = I5,8-2x5,84 = 15.8 to 11.65 = 4,15kp / cm '(a) -6.8 ^o C / 63 ° CP _n -P = 26.3-2x 4.08 = 26.3-8.1 5 = 18.1 5kp / cm ²

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_ 21 _

Due to the construction of the machine, it can be impractical to enter the end plane of the compressor, i.e. axially, inject, as the injection port P then normally in the bore in order to move back a corresponding helix angle of the rotor. If you take the given For example, if the wrap angle of the negative rotor is 200, it means a backward movement of the injection port P according to Fig. 2 by 1.016 mm from the outlet end that the optimal injection site, which is equal to an in a previous step determined angle of envelope <I is equal to 215 from the suction side, plus the distance must be added, by which the return movement occurred times the angle of wrap divided by the diameter of the rotor times the rotor ratio of L / D. Assuming the rotor has a diameter of 102 mm (~ 4 ") and the L / D ratio is 1, results

L ¹ = 0.1016 mm from the outlet end <I = 215 °
(M = wrap angle = 200 °

D = 102 mm = 4 L / D = 1

Hence it follows:

'= 215 + 0.1016 χ) ^{= 215 + 2o} ° ^{= 235} °

To determine the optimal injection position, it is necessary to adapt the built-in pressure ratio to the actual Liche operating pressure ratios are essential. When the actual compression ratio C / R is low

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is variable, there is a greater latitude in the arrangement of the injection site, since on the one hand the for Leakage flow directed towards the suction side and, on the other hand, over-compression Before the outlet expulsion, what is of interest is what the injection site further backwards towards the suction side represents. The best braking performance / t is obviously present when between the position of the outlet opening and an absolute adaptation is given to the actual operating parameters of the compressor, and thus to the injection site for the liquid is not earlier than necessary.

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Claims (5)

Claims 1.J Rotary helical screw compressor for a cooling system with a condenser connected to the outlet side of the compressor, the compressor comprising: a housing which forms a working space comprising two penetrating bores, one in the Suction opening opening into the working space at one end of the housing for supplying a gaseous coolant as the compressor working medium, one opening into the working space Outlet opening in the housing for the discharge of the working medium after it has been compressed, and two in the bores rotatably arranged, intermeshing screw rotors, which, together with the bores, form working chambers, the compressor injecting the compressed refrigerant gas into the Gives condenser of the cooling system and the cooling system has a device to a part of the coolant in liquid Form branch off from the condenser and an injection opening opening into the working space is provided to allow the diverted liquid coolant into the working space for cooling the working medium while it is being compressed inject, characterized in that the injection opening for the liquid coolant is arranged at a point relative to a certain working chamber of the compressor, which is by the current Pressure relationship between the diverted liquid coolant and the pressure of the compressor working medium it is determined during the compression in the chamber, - 24th 509829/0 5 23 so that the cooling of the working medium during the Compression by continuous injection of the diverted liquid coolant at the point independently from an over- or under-compression of the working medium and without significant impairment of the compressor performance is guaranteed. 2. Compressor according to claim 1, characterized in that the injection opening in the the negative screw rotor receiving bore is arranged, said rotor having an envelope angle of 200 ° and the opening in said Bore at a wrap angle from the suction side in the range of 210 ° to 226 °. 3. Compressor according to claim 2, characterized in that g e k e η η - · draws that the injection port for the liquid coolant in terms of the envelope angle of the negative rotor from the suction side 215 ° is arranged. 4. Method of determining the optimal location for the Injection of a liquid coolant into the working space of a rotary screw compressor, the Is part of a cooling system with a condenser coupled to the outlet side of the compressor, the compressor comprising: a housing having a working space comprising two penetrating bores forms a suction opening opening into the work space at one end of the housing to supply a gaseous coolant as a compressor working medium, - 25 509829/0523 an outlet opening in the housing opening into the working space in order to expel the compressed working medium for forwarding to the condenser, and two _{screw rotors which are rotatably arranged in the bores /} meshing with one another and which determine the compressor capacity Releases coolant system, and the system has a device to divert part of the coolant in liquid form from the condenser and inject the diverted liquid coolant via an injection opening into one of the bores for cooling the gaseous working medium during its compression, characterized in that the minimum determines the built-in volume ratio of the compressor, determines within which bore the injection should take place, determines the minimum and maximum operating compression ratio of the compressor, determines the parameters determined with the pressure-envelope angle diagram for the rotor de r brings the selected bore into relation and from the diagram a suitable location for the injection opening in relation to the envelope angle from the suction inlet of the screw rotor for the selected bore is determined, whereby a continuous injection of the liquid coolant into the working space independent of an under- or over-compression of the Working medium results within the limits set by the determined parameters without significantly affecting the compressor performance. - 26 50 9 829/0523 5. The method according to claim 4, characterized in that the injection opening radially and at an axially offset location relative to the outlet end of the bore, wherein the determined injection angle for a variation of the envelope angle of the selected one Corrected the bore of the supported rotor as a result of the axial displacement of the opening according to the following equation will: where mean: = Injection angle after compensation for the rotor envelope angle due to the axial offset of the opening in degrees = uncompensated injection angle in degrees L ¹ = selected axial distance from the outlet end of the working space to the injection opening in mm = Wrap angle of the screw rotor for the selected hole in degrees D = diameter of the screw rotor in mm L / D = ratio of rotor length to diameter. 509829/0523