MANUFACTURING THEREOF

TECHNICAL FIELD [0001] The present disclosure relates generally to compressors; and more specifically, to a single stage rotary screw compressor. Moreover, the present disclosure relates to a method of manufacturing the single stage rotary screw compressor.

BACKGROUND [0002] Compressed air has become a major element of many industries and therefore regarded as a fourth utility after electricity, natural gas and water. The compressed air may be used in various fields, such as manufacturing industries, health industries, entertainment industries, railway industry and so forth. Typically, the compressed air may be generated using a compressor, i.e. an instrument which compresses (i.e. to increase the pressure and decrease the volume) the air. A rotary-screw compressor is an air compressor that comprises a compression-chamber having two rotors skewed at an angle and meshed together. A process of air compression by the rotary-screw compressor can be at a single stage or at a multistage. Further, a compression medium, i.e. air, is fed into the compression-chamber through an inlet, thereafter the air is allowed to pass through the skewed rotors for compression, and finally the compressed air is obtained through an outlet. Typically, the process of air compression in the rotary-screw compressors can be accomplished using dry air and oil infused air.

[0003] However, there are various limitations associated with the conventional air compressors (for example, with rotary-screw compressors), for example primarily related to controlling a temperature of the compressed air to be generated by the compressor. Such conventional air compressors use fluids such as oil, oil based-products, slurries and so forth for controlling the temperature of the compressed air within the air compressors. However, such fluids tend to escape into a compression-chamber of the air compressors, thereby hindering an operation of various moving components (such as the two rotors, shaft coupled with the rotors, and so forth) of the air compressors. Thus, the use of such fluids is generally inefficient at reducing the temperature of the compressed air and further leads to increase in the temperature within the air compressors, such as, due to a friction experienced by the moving components (such as the two rotors). Consequently, the increase in temperature within the air compressors leads to thermal deformation of components (such as housing, compression-chamber and so forth) of the air compressors, thus, causing damage, ineffective operation and reduced operating life thereof.

[0004] Therefore, in light of the foregoing discussion, there exists a need to overcome the aforementioned drawbacks associated with compressors. SUMMARY

[0005] The present disclosure seeks to provide a single stage rotary screw compressor. The present disclosure also seeks to provide a method of manufacturing the single stage rotary screw compressor. The present disclosure seeks to provide a solution to the existing problem of thermal deformation experienced by the conventional air compressors. An aim of the present disclosure is to provide a solution that overcomes at least partially the problems encountered in prior art, and provides efficient, cost effective and reliable single stage rotary screw compressor. [0006] In one aspect, an embodiment of the present disclosure provides a single stage rotary screw compressor comprising :

- a housing having a working-fluid inlet, a compression-chamber, and a working-fluid outlet, wherein the working-fluid inlet and the working-fluid outlet are fluidically coupled to the compression-chamber;

- a rotary-screw mechanism arranged within the compression-chamber, the rotary-screw mechanism comprising a pair of rotors defining an inlet- end and an outlet-end of the rotary-screw mechanism, wherein the rotary-screw mechanism draws the working-fluid into the compression- chamber through the working-fluid inlet, receives the working-fluid between the pair of rotors from the inlet-end, compresses the working- fluid between the pair of rotors, releases the compressed working-fluid from the outlet-end and discharges the compressed working-fluid from the compression-chamber from the working-fluid outlet;

- a volatile-liquid injecting arrangement having at least one volatile-liquid injecting point arranged on the compression-chamber near the outlet- end of the rotary-screw mechanism, wherein the at least one volatile- liquid injecting point provides a specific quantity of volatile-liquid around the outlet-end of the rotary-screw mechanism to allow the specific quantity of volatile-liquid to:

- evaporate and reduce a temperature of the compressed working- fluid released through the outlet-end; and

- evaporate and reduce a temperature of the pair of rotors at the outlet-end, to reduce a thermal deformation experienced by each rotor of the pair of rotors during compression of the working-fluid therebetween;

and

- a cooling jacket arranged on the housing near the outlet-end of the rotary-screw mechanism, wherein the cooling jacket reduces the temperature of the pair of rotors at the outlet-end, to reduce the thermal deformation experienced by each rotor of the pair of rotors. [0007] In another aspect, an embodiment of the present disclosure provides a method of manufacturing a single stage rotary screw compressor, the method comprising :

- providing a housing having an working-fluid inlet, a compression- chamber, and an working-fluid outlet, wherein the working-fluid inlet and the working-fluid outlet are flu id ically coupled to the compression- chamber;

- arranging a rotary-screw mechanism within the compression-chamber, the rotary-screw mechanism comprising a pair of rotors defining an inlet- end and an outlet-end of the rotary-screw mechanism, wherein the rotary-screw mechanism draws working-fluid into the compression- chamber through the working-fluid inlet, receives the working-fluid between the pair of rotors from the inlet-end, compresses the working- fluid between the pair of rotors, releases the compressed working-fluid from the outlet-end and discharges the compressed working-fluid from the compression-chamber from the working-fluid outlet;

- providing a volatile-liquid injecting arrangement having at least one volatile-liquid injecting point arranged on the compression-chamber near the outlet-end of the rotary-screw mechanism, wherein the at least one volatile-liquid injecting point provides a specific quantity of volatile-liquid around the outlet-end of the rotary-screw mechanism to allow the specific quantity of volatile-liquid to:

- evaporate and reduce a temperature of the compressed working- fluid released through the outlet-end; and

- evaporate and reduce a temperature of the pair of rotors at the outlet-end, to reduce a thermal deformation experienced by each rotor of the pair of rotors during compression of the working-fluid therebetween;

and

- arranging a cooling jacket on the housing near the outlet-end of the rotary-screw mechanism, wherein the cooling jacket reduces the temperature of the pair of rotors at the outlet-end, to reduce the thermal deformation experienced by each rotor of the pair of rotors.

[0008] Embodiments of the present disclosure substantially eliminate or at least partially address the aforementioned problems in the prior art, and enable controlling of a temperature of the compressed working-fluid to be generated by the single stage rotary screw compressor. Moreover, temperature uniformity of components (such as housing, compression-chamber and so forth) of the single stage rotary screw compressor is achieved. Furthermore, reduction in the temperature of rotating components (such as the pair of rotors) of the compressors during compression of the working-fluid is obtained. Therefore, the present disclosure provides the single stage rotary screw compressor that overcomes the problems associated with high temperatures in the compressor, and enables efficient, cost effective and reliable operation of the single stage rotary screw compressor.

[0009] Additional aspects, advantages, features and objects of the present disclosure would be made apparent from the drawings and the detailed description of the illustrative embodiments construed in conjunction with the appended claims that follow. [00010] It will be appreciated that features of the present disclosure are susceptible to being combined in various combinations without departing from the scope of the present disclosure as defined by the appended claims.

BRIEF DESCRIPTION OF THE DRAWINGS [00011] The summary above, as well as the following detailed description of illustrative embodiments, is better understood when read in conjunction with the appended drawings. For the purpose of illustrating the present disclosure, exemplary constructions of the disclosure are shown in the drawings. Flowever, the present disclosure is not limited to specific methods and instrumentalities disclosed herein. Moreover, those skilled in the art will understand that the drawings are not to scale. Wherever possible, like elements have been indicated by identical numbers.

[00012] Embodiments of the present disclosure will now be described, by way of example only, with reference to the following diag rams wherein :

FIG. 1 is a block d iagram of a single stage rotary screw compressor, in accordance with an embodiment of the present disclosure;

FIG. 2 is a simulated perspective view of the single stage rotary screw compressor of the FIG. 1, in accordance with an embodiment of the present disclosure;

FIG. 3 is a front view of a rotor arrangement of the single stage rotary screw compressor of FIGs. 1 or 2, in accordance with an embodiment of the present disclosure;

FIG. 4 is a schematic illustration of blocking on a female rotor profile at its outer circle, in accordance with an embodiment of the present d isclosure; FIG. 5 is a schematic illustration of a rack curve configuration between the rotors, in accordance with an embodiment of the present d isclosure;

FIG. 6 is a schematic illustration depicting interaction between the main and female rotors profiles, in accordance with an embodiment of the present disclosure;

FIG. 7 is a tabular representation of experimental data to observe operating parameters of the single stage rotary screw compressor (of FIGs. 1-3), in accordance with an embodiment of the present disclosure;

FIG. 8 is a graphical representation of pressure variation in the compression-chamber for the experimented cases of FIG. 7, in accordance with an embodiment of the present disclosure;

FIG. 9 is a graphical representation of torque variation in the compression-chamber for the experimented cases of FIG. 7, in accordance with an embodiment of the present disclosure;

FIG. 10 is a graphical representation of power variation in the compression cycle of the compression-chamber for the experimented cases of FIG. 7, in accordance with an embodiment of the present disclosure;

FIG. 11 is a graphical representation of power variation in the interference and the magnitude of the rotors during the experimented cases of FIG. 7, in accordance with an embodiment of the present disclosure;

FIG. 12 is a graphical representation of power variation in the interference of the rotor profiles during the experimented cases of FIG. 7, in accordance with an embodiment of the present disclosure; FIG. 13 is a graphical representation of experimented cases for observing volatile-liquid mass requirement on varying a compression power of the single stage rotary screw compressor, in accordance with an embodiment of the present disclosure;

FIG. 14 is a graphical representation of experimented cases for observing delivery temperature on varying compression power of the single stage rotary screw compressor, in accordance with an embodiment of the present disclosure; FIG. 15 is a graphical representation of experimented cases for observing relative volatile-liquid mass requirement on varying the compression power of the single stage rotary screw compressor, in accordance with an embodiment of the present disclosure; and

FIG. 16 is a graphical representation of experimented cases for observing delivery temperature on varying the compression power of the single stage rotary screw compressor, in accordance with an embodiment of the present disclosure. [00013] In the accompanying drawings, an underlined number is employed to represent an item over which the underlined number is positioned or an item to which the underlined number is adjacent. A non- underlined number relates to an item identified by a line linking the non- underlined number to the item. When a number is non-underlined and accompanied by an associated arrow, the non-underlined number is used to identify a general item at which the arrow is pointing.

DETAILED DESCRIPTION OF EMBODIMENTS

[00014] The following detailed description illustrates embodiments of the present disclosure and ways in which they can be implemented. Although some modes of carrying out the present disclosure have been disclosed, those skilled in the art would recognize that other embodiments for carrying out or practicing the present disclosure are also possible.

In one aspect, the present disclosure provides a single stage rotary screw compressor comprising :

- a housing having a working-fluid inlet, a compression-chamber, and a working-fluid outlet, wherein the working-fluid inlet and the working-fluid outlet are fluidically coupled to the compression-chamber;

- a rotary-screw mechanism arranged within the compression-chamber, the rotary-screw mechanism comprising a pair of rotors defining an inlet- end and an outlet-end of the rotary-screw mechanism, wherein the rotary-screw mechanism draws the working-fluid into the compression- chamber through the working-fluid inlet, receives the working-fluid between the pair of rotors from the inlet-end, compresses the working- fluid between the pair of rotors, releases the compressed working-fluid from the outlet-end and discharges the compressed working-fluid from the compression-chamber from the working-fluid outlet;

- a volatile-liquid injecting arrangement having at least one volatile-liquid injecting point arranged on the compression-chamber near the outlet- end of the rotary-screw mechanism, wherein the at least one volatile- liquid injecting point provides a specific quantity of volatile-liquid around the outlet-end of the rotary-screw mechanism to allow the specific quantity of volatile-liquid to:

- evaporate and reduce a temperature of the compressed working- fluid released through the outlet-end; and

- evaporate and reduce a temperature of the pair of rotors at the outlet-end, to reduce a thermal deformation experienced by each rotor of the pair of rotors during compression of the working-fluid therebetween;

and [00015] - a cooling jacket arranged on the housing near the outlet- end of the rotary-screw mechanism, wherein the cooling jacket reduces the temperature of the pair of rotors at the outlet-end, to reduce the thermal deformation experienced by each rotor of the pair of rotors.

In another aspect, the present disclosure provides a method of manufacturing a single stage rotary screw compressor, the method comprising : - providing a housing having an working-fluid inlet, a compression- chamber, and an working-fluid outlet, wherein the working-fluid inlet and the working-fluid outlet are flu id ically coupled to the compression- chamber;

- arranging a rotary-screw mechanism within the compression-chamber, the rotary-screw mechanism comprising a pair of rotors defining an inlet- end and an outlet-end of the rotary-screw mechanism, wherein the rotary-screw mechanism draws the working-fluid into the compression- chamber through the working-fluid inlet, receives the working-fluid between the pair of rotors from the inlet-end, compresses the working- fluid between the pair of rotors, releases the compressed working-fluid from the outlet-end and discharges the compressed working-fluid from the compression-chamber from the working-fluid outlet;

- providing a volatile-liquid injecting arrangement having at least one volatile-liquid injecting point arranged on the compression-chamber near the outlet-end of the rotary-screw mechanism, wherein the at least one volatile-liquid injecting point provides a specific quantity of volatile-liquid around the outlet-end of the rotary-screw mechanism to allow the specific quantity of volatile-liquid to:

- evaporate and reduce a temperature of the compressed working- fluid released through the outlet-end; and

- evaporate and reduce a temperature of the pair of rotors at the outlet-end, to reduce a thermal deformation experienced by each rotor of the pair of rotors during compression of the working-fluid therebetween;

and

[00016] - arranging a cooling jacket on the housing near the outlet- end of the rotary-screw mechanism, wherein the cooling jacket reduces the temperature of the pair of rotors at the outlet-end, to reduce the thermal deformation experienced by each rotor of the pair of rotors. [00017] Referring to FIG. 1, illustrated is a block diagram of a single stage rotary screw compressor 100, such as a twin rotary-screw compressor, in accordance with an embodiment of the present disclosure. As shown, the single stage rotary screw compressor 100 comprises a housing 102 having compression-chamber 104, a working-fluid inlet 106 and a working-fluid outlet 108. The working-fluid inlet 106 and the working-fluid outlet 108 are fluidically coupled to the compression- chamber 104. The single stage rotary screw compressor 100 also comprises a rotary-screw mechanism 110 arranged within the compression-chamber 104. The rotary-screw mechanism 110 comprises a pair of rotors. As shown, the rotary-screw mechanism 110 comprises a main rotor 112 and a female rotor 114 configured to mesh with each other (profiles of the main rotor 112 and the female rotor 114, are best shown in FIGs. 3-6). Furthermore, each of the main rotor 112 and the female rotor 114 includes an inlet-end, such as an inlet-end 116 in proximity to the working-fluid inlet 106, and an outlet-end, such as an outlet-end 118 in proximity to the working-fluid outlet 108. The rotary- screw mechanism 110 draws working-fluid into the compression- chamber 104 through the working-fluid inlet 106, receives the working- fluid between the pair of rotors from the inlet-end 116, compresses the working-fluid between the pair of rotors, releases the compressed working-fluid from the outlet-end 118 and discharges the compressed working-fluid from the compression-chamber 104 from the working-fluid outlet 108. The working-fluid can include fluids such as air, nitrogen, carbon dioxide, inert gases, vapors and so forth. The working-fluid in an uncompressed state is drawn into the single stage rotary screw compressor 100 by the rotary-screw mechanism 110 through the working-fluid inlet 106 that is fluidically coupled to the compression- chamber 104. It will be appreciated that rotation of the main rotor 112 and the female rotor 114 creates suction therebetween and consequently, the working-fluid in the uncompressed state that is drawn through the working-fluid inlet 106 is received between the pair of rotors. Thereafter, the rotation of the main rotor 112 and the female rotor 114 enables compression of the working-fluid in the uncompressed state, such as, due to reduction of intermolecular space between molecules of the working-fluid due to mutual rotation of the pair of rotors. Furthermore, such a compression of the working-fluid increases a temperature thereof. Moreover, the rotation of the main rotor 112 and the female rotor 114 pushes the compressed working-fluid towards the outlet-end 118 of the rotary-screw mechanism 110 wherefrom, the compressed working-fluid is released out of the compression-chamber 104. Subsequently, the compressed working-fluid is discharged from the compression-chamber 104 from the working-fluid outlet 108 that is fluidically coupled to the compression-chamber 104. The rotary-screw mechanism 110 also includes a shaft 120 coupled to the main rotor 112 and configured to provide a rotary motion (for example from an engine) to the main rotor 112 for the rotation of the main rotor 112. The main rotor 112 further provides rotary motion to the female rotor 114 for the rotation thereof. The rotary-screw mechanism 110 is operable to draw working-fluid into the compression-chamber 104 through the working- fluid inlet 106, compress the drawn working-fluid within the compression-chamber 104 and release the compressed working-fluid through the working-fluid outlet 108.

[00018] Referring now to FIG. 2, illustrated is a simulated perspective view of the single stage rotary screw compressor 100 of FIG. 1, in accordance with an embodiment of the present disclosure. As shown, the FIG. 2 primarily depicts the working-fluid inlet 106, the working-fluid outlet 108, the main rotor 112, the female rotor 114, and the at least one volatile-liquid injecting point (such as a volatile-liquid injecting point 140 arranged on one side of the compression-chamber 104 and another volatile-liquid injecting point 142 arranged on an opposite side of the compression-chamber 104). The volatile-liquid can include liquids such as water, alcohol-based coolants, calcium chloride solution, refrigerants (for example, R-134a) and so forth. It will be appreciated that the volatile-liquids include liquids that evaporate at a low temperature, thereby, enabling extraction of heat from another substance (such as, the pair of rotors) due to latent heat of vaporization absorbed by the volatile-liquids. Consequently, upon absorption of the heat of vaporization, the volatile-liquids undergo a change in phase from liquid- phase to vapor-phase.

[00019] Referring now to FIG. 3, illustrated is a front view of the rotary-screw mechanism 110 (i.e. the main rotor 112 and the female rotor 114) of the single stage rotary screw compressor 100 of FIGs. 1 or 2, in accordance with an embodiment of the present disclosure. Specifically, FIG. 3 illustrates profiles of the main rotor 112 and the female rotor 114.

[00020] According to an embodiment, a separation between the rotors (such as the main rotor 112 and the female rotor 114) of the pair of rotors is in a range of 10 microns to 50 microns. Such a gap is selected based on a required operating temperature between the rotors (112, 114) during operation of the single stage rotary screw compressor 100. The main rotor 112 and the female rotor 114 are designed such that during rotation thereof through a rotation cycle, an interlobe clearance between the main rotor 112 and the female rotor 114 is reduced to a minimum value (for example, in a range of 10 microns to 30 microns) under operating temperature that lies between temperature of the single stage rotary screw compressor 100 at cold start, to temperature of single stage rotary screw compressor 100 when operating at working-fluid pressure of 11 bar or more. For example, the main rotor 112 and the female rotor 114, i.e. outer surfaces thereof are spaced apart by a distance of 30 microns. The design of main rotor 112 and the female rotor 114 can be designed using SCORPATFI design software. The main rotor 112 and the female rotor 114 can be designed such that the rotors only have three points of contact therebetween during operation of the single stage rotary screw compressor 100, thus, minimizing substantial metal-to-metal contact and further reducing operating temperature due to friction created between the rotors. In an example, the interlobe clearance between the main rotor 112 and the female rotor 114 is less than or equal to 30 microns depending on angular positions thereof respectively.

[00021] Also, a separation between the pair of rotors and the compression-chamber 104 is in a range of 10 microns to 50 microns. For example, a gap size of 30 microns is there between the rotors (i.e. the main rotor 112 and the female rotor 114) and the housing 102 (particularly a housing bore).

[00022] Referring now to FIG. 4, illustrated is a schematic illustration of blocking on the female rotor 114 profile at its outer circle, in accordance with an embodiment of the present disclosure. Blocking relates to uniform distribution of nodes on rotor profile and on outer circle. Furthermore, the blocking allows a block to be less refined compared to the final grid. Moreover, the blocking allows the block to be used as a reference for refinement in any required regions. Furthermore, the blocking allows the block to be calculated only once and thereby, rotate for various rotor positions. Points distributed on boundaries are represented in index notation with respect to the physical coordinate system as r (x, y). Furthermore, points on the rotor profile (i.e. female rotor 114) are n,j=o (x, y) and points on outer boundary consisting of casing and rack curve are n,j=i (x,y) and the point distribution on outer full circle is n,j'=i.(x, y). Each background blocks is identified by its index Bi. Further, the points on the inner boundary of the blocks which are the rotor profile nodes are r bi,j=o (x, y) and point distribution on outer full circle is r bi,j _=i .(x, y). The nodes are distributed on the outer circle covering the rack part with required number of points irack. Further, the nodes are distributed on the outer circle covering the casing part with required number of points l _easing . Subsequently, the nodes data are available for r bi ₌ o (x, y), r bi,j _=i .(x, y) and n,j' _=i .(x, y) is required to calculate n,j ₌ o (x,y). The node distribution is based on equidistant spacing as given in equation below: n,j' _=i (x,y)= n _=i,j'=i (x, y) + Sii wherein, Si = Si/I and Si = n,j' _=i . (x,y) - n _=0j'=i . (x,y)

[00023] A scanning function has the information of the background blocking. The scanning function traces each node n, _{j = i} . (x,y) and identifies the block B, to which node belongs. In a case, a single block can have multiple nodes or there can be blocks with no nodes, as the distribution on the rack curve can be refined in comparison to the blocking. Similarly, distribution on casing can be coarse in comparison to the blocking. Further, once the nodes associated with each block are traced by the scanning function, an arc-length based projection is used to determine the nodes n,j ₌ o (x,y) to be placed on rotor profile (i.e. female rotor 114). Subsequently, time constraint is imposed on the node placement to bound in the same block B,. Further, B, as that of the outer circle nodes n,j' _=i (x,y). Further, the illustration provides projection of ^r U' _=i (x,y) on the inner boundary of the block to get n,j ₌ o (x,y), wherein the projection is based on arc length factor given by equation below: n,j ₌ o (x,y) = r bi,j ₌ o (x,y) + (r b _i+i,j= o (x,y) - r bij ₌ o(x,y))Si/Si

wherein, S, = h,^ ₌ i (x,y) - r bi, _{j =i} (x,y)

Si = r bi _+i,j'=i (x, y) - r by-i (x,y)

Further, the calculated positions n,j ₌ o (x,y) of nodes ensure that they are guided by a regular rotor profile. Further, regularized distribution is superimposed onto the rack curve by finding the intersection points of the distribution lines and the rack curve.

[00024] Referring to FIG. 5, illustrated is a schematic illustration of a rack curve configuration between the rotors (i.e. the main rotor 112 and the female rotor 114), in accordance with an embodiment of the present disclosure. For example, FIG. 5 depicts that the intersection points are the new distributions n,j = i (x,y) on the rack curve. Further, as blocks on main rotor 112 side are different from blocks on female rotor 114 side, the intersection points obtained on the common rack curve from the two blocks can be identical or non-identical. Depending on the obtained intersection points a conformal or non-conformal map between the two rotor blocks is further obtained. Further, with the blocking approach the 3D grid is fully hexahedral and both the main and female rotors 112, 114 surfaces are smoothly captured. [00025] FIG. 6 is a schematic illustration depicting interaction between the main and female rotors 112, 114 profiles, in accordance with an embodiment of the present disclosure. Small non-aligned node movements are possible at the transition point, from interlobe region to the casing region. Flowever, these are positioned on the surface of rotors (i.e. the main and female rotors 112, 114) and do not result in any irregular cells. Further, surface mesh on the casing is of the highest quality with regular quadrilateral cells. Moreover, the surface mesh on the interlobe interface mostly follows axial grid lines with only small transverse movements in vicinity of top and bottom CUSP's which are cyclically repeating, wherein the movements are on the surface of the interface and do not result in any irregular cells. Therefore, the current implementation allows for a fully conformal interface with the equal index of the top and bottom CUSP points which ensures straight line in the axial direction. In order to design the single stage rotary screw compressor 100 (of the present disclosure) to regulate the temperature of the compressed working-fluid, a simulation of the single stage rotary screw compressor 100 is modelled using software such as ANSYS for Computational Fluid Dynamics (CFD) model, ANSYS Finite Element Analysis (FEA), SCORG grid generator. Particularly, the ANSYS FEA is employed for the thermal deformation analysis of the single stage rotary screw compressor 100. The inputs required for this model are the distribution of temperature on boundaries of the single stage rotary screw compressor components. Since the CFD model calculates only flow domains (for example, such as flow of working-fluid) of the single stage rotary screw compressor, no solid components are included in the CFD model. The temperature results obtained from the CFD model are available for the flow domain but temperature of a surface of the rotor, compressor chamber and other components are also required. Therefore, the thermal deformation model has three main modules, namely transfer of boundary temperatures from CFD to Thermal model, solving thermal conduction equation using FEA in the solid components, solving deformation equation using FEA in the solid components. Moreover, SCORG grid generator is used to generate a numerical grid for rotor domain (such as the main and female rotors 112, 114). Furthermore, the simulation of the single stage rotary screw compressor 100 is utilized to obtain associated operating parameters. Such operating parameters are analyzed and are discussed in conjunction with FIG. 7, discussed herein later.

[00026] In an embodiment, the housing 102 and the pair of rotors are manufactured using Invar, steel, cast iron, iron alloy, aluminum alloy or a combination thereof and the material is heat-treated and/or coated using a different material. Materials such as Invar, steel, cast iron, iron alloy, aluminum alloy or a combination thereof have low thermal expansion coefficients, therefore, such materials resist thermal expansion for a wide range of temperature. Subsequently, the housing 102 and the pair of rotors (i.e. the main rotor 112 and the female rotor 114) manufactured using such materials are able to withstand high temperatures, thereby reducing the thermal deformations experienced by the housing 102 and the pair of rotors (i.e. the main rotor 112 and the female rotor 114). Furthermore, the housing 102 and the pair of rotors can be manufactured using a heat-treated material, such as heat- treated steel. Moreover, the housing 102 and the pair of rotors can be coated with a different material that the material used to manufacture housing 102 and the pair of rotors, such as, the housing 102 and the pair of rotors are manufactured using steel and are coated with zinc. [00027] In an embodiment, the single stage rotary screw compressor

102 further comprises a bearing arrangement arranged within the housing 102. It will be appreciated that the housing 102 may include pockets (or cavities) in order to support the bearing arrangement therein. Specifically, the bearing arrangement comprises a pair of first-bearings 130 to rotatably support the inlet-end 116 of each of the pair of rotors

(i.e. the main rotor 112 and the female rotor 114), and a pair of second- bearings 132 to rotatably support the outlet-end 118 of each of the pair of rotors (i.e. the main rotor 112 and the female rotor 114). It will be appreciated that each rotor of the pair of rotors, such as, the main rotor 112 and the female rotor 114 are supported on shafts that rotate to enable the rotation of the pair of rotors. Furthermore, the shafts supporting each rotor of the pair of rotors extend out of the compression- chamber 104. Thus, the pair of first-bearings 130 support the shafts towards the inlet-end 116 of each of the pair of rotors, such as, a portion of the shafts extending out of the compression-chamber 104 towards the working-fluid inlet 106. The pair of second-bearings 132 support the shafts towards the outlet-end 118 of each of the pair of rotors, such as, a portion of the shafts extending out of the compression-chamber 104 towards the working-fluid outlet 108. [00028] The single stage rotary screw compressor 100 further comprises a volatile-liquid injecting arrangement having at least one volatile-liquid injecting point arranged on a portion 144 of the compression-chamber 104 near the outlet-end 118 of the rotary-screw mechanism 110. As shown, the volatile-liquid injecting arrangement includes the at least one volatile-liquid injecting point 140, 142 (or channel) on a portion 144 of the compression-chamber 104 near the outlet-ends 118 of the rotary-screw mechanism 110 (i.e. the main rotor 112 and the female rotor 114). [00029] The at least one volatile-liquid injecting point 140 and 142 provides a specific quantity of volatile-liquid around the outlet-end 118 of the rotary-screw mechanism 110 (i.e. the main rotor 112 and the female rotor 114) to allow the specific quantity of volatile-liquid to evaporate and reduce a temperature of the compressed working-fluid released through the outlet-end 118. It will be appreciated that as the volatile-liquid having a lower temperature as compared to the temperature of the compressed working-fluid, is allowed to come in contact with the compressed working-fluid, the volatile-liquid derives latent heat of evaporation from the compressed working-fluid in the compression-chamber 104 and evaporates. Thus, as the volatile-liquid injected into the compression-chamber 104 evaporates, the temperature of the compressed working-fluid in the compression-chamber 104 is reduced (under constant pressure conditions). Moreover, the at least one volatile-liquid injecting point provides a specific quantity of volatile-liquid to allow the specific quantity of volatile-liquid to evaporate and reduce a temperature of the pair of rotors (i.e. the main rotor 112 and the female rotor 114) at the outlet-end 118, to reduce a thermal deformation experienced by each rotor of the pair of rotors (i.e. the main rotor 112 and the female rotor 114) during compression of the working-fluid therebetween. It may be appreciated that the volatile-liquid injecting arrangement may further include a volatile-liquid source, pipes (or conduits) coupled to the at least one volatile-liquid injecting point 140, 142 and a pump (not shown), which may be instructed for pumping or injecting the specific quantity of volatile-liquid into the compression- chamber 104 through the at least one volatile-liquid injecting point 140, 142.

[00030] In an embodiment, the at least one volatile-liquid injecting point 140 and 142 provides the specific quantity of volatile-liquid to reduce a temperature of the compression-chamber 104. The at least one volatile-liquid injecting point 140 and 142 injects specific quantity of volatile-liquid into the compression-chamber 104 such that the specific quantity of volatile-liquid evaporates and reduces the temperature of the compression-chamber 104. Notably, reduction in temperature of the compression-chamber 104 reduces the temperature of the pair of rotors (i.e. the main rotor 112 and the female rotor 114). [00031] In an embodiment, the specific quantity of volatile-liquid is less than or equal to a quantity of volatile-liquid required for saturation of the compressed working-fluid at the working-fluid outlet 108. The specific quantity of volatile-liquid is determined such that the specific quantity of volatile-liquid evaporates and saturates the compressed working-fluid at the working-fluid outlet 108. It will be appreciated that the specific quantity of volatile-liquid is required to be less than or equal to the quantity of volatile-liquid required for saturation, as if the specific quantity of volatile-liquid is more than the quantity of volatile-liquid required for saturation of the compressed working-fluid, the specific quantity of volatile-liquid does not evaporate and remains in the single stage rotary screw compressor 100.

[00032] Moreover, the single stage rotary screw compressor 100 further comprises a cooling jacket 150 arranged on the housing 102 (such as on a portion 152 of the housing 102) near the outlet-end 118 of the rotary-screw mechanism 110. The cooling jacket 150 acts as a heat-exchanger for reducing the temperature of various components of the single stage rotary screw compressor 100 such as the pair of rotors at the outlet-end 118, to reduce the thermal deformation experienced by each rotor of the pair of rotors (i.e. the main rotor 112 and the female rotor 114).

[00033] In an embodiment, the cooling jacket 150 is arranged on the housing 102 to enclose the pair of second-bearings 132. For example, the cooling jacket 150 may be hollow housing, arranged on the portion 152 of the housing 102, that receives a coolant, such as volatile-liquid and the like, for taking away heat from the portion 152 of the housing 102 and an area around the compression-chamber 104 in proximity to the working-fluid outlet 108. Optionally, the cooling jacket 150 may include a plate exposed to the ambient environment to conduct heat away from the portion 152 of the housing 102. In such instance, the plate may be a thermally conductive structure and may include a plurality of fins configured on the conductive plate.

[00034] In an embodiment, the single stage rotary screw compressor 100 further comprises a non-return valve 160 flu id ically coupled with the working-fluid outlet 108, for reducing the temperature of the compressed working-fluid at the working-fluid outlet 108 by regulating a pressure thereof. The non-return valve 160 facilitates in reduction of temperature for the compressed working-fluid exiting the working-fluid outlet 108. For example, the non-return valve 160 prevents pressure from returning into the compression-chamber 104 and exhausting through the working- fluid outlet 108 during shut-down. Furthermore, the non-return valve 160, when combined with operation of the main rotor 112 and the female rotor 114 (i.e. rotor lobe compression cycle), produces a pulsation effect which is less than the pressure of the compressed working-fluid exiting the working-fluid outlet 108 for a certain time incessant. This results in a degree of pressure relief in the working-fluid outlet 108 and consequently a reduced temperature thereof.

[00035] In an embodiment, the single stage rotary screw compressor 100 further comprises an working-fluid dryer 170 flu id ically coupled with the non-return valve 160, wherein the working-fluid dryer 170 removes evaporated volatile-liquid from the compressed working-fluid discharged through the working-fluid outlet 108 via the non-return valve 160. It will be appreciated that the process of working-fluid compression concentrates atmospheric contaminants, including water vapor. This raises the dew point of the compressed working-fluid relative to free atmospheric working-fluid and leads to condensation within the working- fluid outlet 108 and the compressed working-fluid cools downstream of the compressor 100.

[00036] Referring to FIG. 7, illustrated is a tabular representation of experimental data to observe operating parameters of the single stage rotary screw compressor 100, in accordance with an embodiment of the present disclosure. As shown, the table 700 represents the variation in mass of volatile-liquid required for saturation of the compressed working- fluid with the variation in the supplied power. In an example, the operating parameters are observed when the required pressure of the compressed working-fluid is 11 bars. As shown, the table depicts various cases, namely case 1, case 2, case 3 and case 4, at various speeds, pressures, volatile-liquid masses, and saturation masses and average discharged temperatures. In an example, considering low volatile-liquid mass flow rate of 0.009 Kg/sec the cooling effect is stronger in Case 2 at 4500 rpm compared to Case 1 at 6000rpm which has 2x volatile-liquid mass flow compared to Case 2. In another example, case 3 and case 4 are designed such that the mass flow rate of volatile-liquid is 5x and lOx of the saturation mass of Case 2 respectively with the aim of achieving a discharge temperature lower than 200°C. In yet another example, in case 3, a temperature of 205°C is achieved at 4500 rpm and in case 4, a temperature of 187°C is achieved at 6000rpm with the increased mass flow of volatile-liquid.

[00037] Referring to FIG. 8, illustrated is a graphical representation of pressure variation in the compression-chamber 104 for the experimented cases of FIG. 15, in accordance with an embodiment of the present disclosure. Specifically, FIG. 8 is a graph 800 that describes the pressure variation in the compression-chamber 104. In an example, the graph 800 depicts variation in the pressure, such that the Y axis of the graph 800 depicts amount of pressure in the compression-chamber 104 when the main rotor 112 is operational and X axis of the graph 800 depicts the angle at which main rotor 112 may be arranged. Furthermore, for an ideal condition wherein the main rotor 112 is operating at a fixed revolution per minute of 6000 rpm, and with a fixed volatile-liquid mass flow rate is 0.018 Kg/sec (as described in the case 1 of FIG. 7), a strong pressure pulse is generated in the working-fluid outlet 108 which is depicted by showing an elevation in the graph. Additionally, as shown, the ideal condition may generate a high compression resulting into a steep pressure rise at 350-degree of the rotor angle. It will be appreciated that, if the pressure variation is calculated at 11.0 bar discharge pressure the cases 2, 3 and 4 of FIG. 7 can be described using a similar graph as shown herein.

[00038] Referring to FIG. 9, illustrated is a graphical representation of torque variation in the compression-chamber for the experimented cases of FIG. 7, in accordance with an embodiment of the present disclosure. Specifically, FIG. 9 is a graph 900 that describes the torque variation in the compression-chamber 104. In an example, the graph 900 depicts variation in the torque of the two rotors, namely the main rotor 112 and the female rotor 114, in two compression cycles, such that the Y axis of the graph 900 depicts amount of torque in the compression- chamber 104 when the main and the female rotors 112, 114 is operational and X axis of the graph 900 depicts the angle at which the main and the female rotors 112, 114 may be arranged. Furthermore, for an ideal condition wherein the main rotor 112 and the female rotor 114 are operating at a fixed revolution per minute of 6000 rpm, and with a fixed volatile-liquid mass flow rate is 0.018 Kg/sec (as described in the case 1 of FIG. 7), an average torque of 30.0 Nm (approximate) for the main rotor 112 and an average torque of 3.69 Nm (approximate) for the female rotor 114 may be achieved. Optionally, the female rotor 114 operates in an opposite direction to that of the main rotor 112. Furthermore, it will be appreciated that the if the variation in the torque of the two rotors 112, 114 is calculated at 11.0 bar discharge pressure, the resultant rotor torque for the cases 2, 3 and 4 of FIG. 7 can be described using a similar graph as shown herein. [00039] Referring to FIG. 10, illustrated is a graphical representation of power variation in the compression cycle of the compression-chamber for the experimented cases of FIG. 7, in accordance with an embodiment of the present disclosure. Specifically, FIG. 10 is a graph lOOO that describes the power variation in the compression cycle. In an example, the graph 1000 depicts variation in the power, such that the Y axis of the graph 1000 depicts amount of power during the compression cycle when the main rotor 112 is operational, and X axis of the graph 1000 depicts the angle at which main rotor 112 may be arranged. Furthermore, for an ideal condition wherein the main rotor 112 is operating at a fixed revolution per minute of 6000 rpm, and with a fixed volatile-liquid mass flow rate is 0.018 Kg/sec (as described in the case 1 of FIG. 7), an average power may be 21.0 kW. Additionally, for another condition wherein the main rotor 112 is operating at a fixed revolution per minute of 4500 rpm, and with a fixed volatile-liquid mass flow rate is 0. 009 Kg/sec (as described in the case 2 of FIG. 7), an average power may be 15.0 kW. [00040] Referring to FIG. 11, illustrated is a graphical representation of power variation in the interference and the magnitude of the rotors during the experimented cases of FIG. 7, in accordance with an embodiment of the present disclosure. Specifically, FIG. 11 is a graph 1100 that depicts the superimposition of a deformation helixes of main rotor's 112 root and female rotor's 114 tip. In an example, the graph 1100 depicts the superimposition, such that the Y axis of the graph 1100 depicts the deformation reading for the main and female rotors 112, 114 and X axis of the graph 1100 depicts the length for normalized helix. Furthermore, the graph 1100 represents the interference of the rotors 112, 114 during operation and the magnitude of the deformation indicates a likely seizure of the rotors. It will be appreciated that, the graph 1100 is formed based on the gap sizes of 30 microns between the rotors 112, 114, and between the rotors 112, 114 and housing 102 (particularly, housing bore).

[00041] Referring to FIG. 12, illustrated is a graphical representation of power variation in the interference of the rotor profiles during the experimented cases of FIG. 7, in accordance with an embodiment of the present disclosure. Specifically, FIG. 12 is a graph 1200 that depicts temperature distribution and resultant thermal deformation of the main rotor's 112 lobe profile at discharge end of the single stage rotary screw compressor (such as the single stage rotary screw compressor 100 of the FIG. 1). Optionally, the compression of the working-fluid in the single stage rotary screw compressor 100 causes the temperature difference, thereby causing a resultant thermal deformation. As shown in the graph 1200 the deformed profiles have been plotted on a normalized length scale representing the interference of the rotor profiles during operation.

[00042] Referring to FIG. 13, illustrated is a graphical representation of experimented cases for observing volatile-liquid mass requirement on varying the compression power of the single stage rotary screw compressor 100, in accordance with an embodiment of the present disclosure. Specifically, FIG. 13 is a graph 1300 that represents the relation between the volatile-liquid mass requirement with respect to the compression power of the single stage rotary screw compressor 100. Furthermore, the observations have been carried along with an increase in the suction temperature (from 10°C to 25°C). The graph 1300 shows that there is a proportional increment in volatile-liquid mass requirement from 0.2 kg/min at lOkW power to 1.2 kg/min at 60kW. Moreover, on increasing the suction temperature from 10°C to 25°C, resulted in only a small increase in volatile-liquid mass requirement.

[00043] Referring to FIG. 14, illustrated is a graphical representation of experimented cases for observing delivery temperature on varying compression power of the single stage rotary screw compressor 100, in accordance with an embodiment of the present disclosure. As shown, FIG. 14 is graph 1400 that represents the relation between the delivery temperature with respect to the compression power of the single stage rotary screw compressor 100. Furthermore, the observations have been carried along with an increase in the suction temperature (from 10°C to 25°C). The graph 1400 shows that on varying the compression power from 10 kW to 60 kW, the delivery temperature with saturation condition can vary from 100°C to 150 °C. Moreover, increasing suction temperature from 10°C to 25°C resulted in very close delivery temperatures as compensated by an incremental evaporated volatile- liquid mass.

[00044] Referring to FIG. 15, illustrated is a graphical representation of experimented cases for observing relative volatile-liquid mass requirement on varying the compression power of the single stage rotary screw compressor 100, in accordance with an embodiment of the present disclosure. As shown, FIG. 15 is a graph 1500 that represents the relation between the relative volatile-liquid mass with respect to the compression power of the single stage rotary screw compressor 100. The graph 1500 shows that on varying the compression power from 10 kW to 60 kW, at 10 kW power a 2%, 4% and 8% higher volatile-liquid mass is required to produce saturation with 15% RH (relative humidity), 25°C -3% RH (relative humidity) and 25°C-15% RH (relative humidity) suction condition respectively. Also, the incremental value drops to a very low percentage is less than 1% at 60 kW compression power.

[00045] Referring to FIG. 16, illustrated is a graphical representation of experimented cases for observing delivery temperature on varying the compression power of the single stage rotary screw compressor 100, in accordance with an embodiment of the present disclosure. As shown, FIG. 16 is a graph 1600 that represents the relation between the delivery temperature with respect to the compression power of the single stage rotary screw compressor 100. The graph 1600 shows that there is no difference in delivery temperature when the suction humidity is increased. Also, the intake relative humidity differences can be compensated to produce the same delivery temperature by varying the evaporated volatile-liquid mass.

[00046] The results of the experiments conducted for calculating the requirement of volatile-liquid mass to achieve saturation at working-fluid outlet 108 shows that the volatile-liquid mass of 0.009 Kg/sec is required to achieve saturated working-fluid at the working-fluid outlet 108 with power dissipation of 30 kW approximately. However, the experiments do not account for transient and leakage affects. Additionally, CFD calculation has therefore resulted in higher than saturation exit temperatures.

[00047] The experiments for calculating the working-fluid temperature distribution inside the single stage rotary screw compressor 100 is conducted. The experiment is conducted taking an iso-surface generated with liquid volume fraction of 0.0001. Furthermore, the result shows that the temperature in the working-fluid inlet 106 is lower on the female rotor 114 side, but on the main rotor 112 side shows higher working-fluid temperature. This indicates that the leakage is higher from the tip of the main rotor 112 as compared to the female rotor 114, also that the cooling is more effective on the female rotor 114 side as compared to the main rotor 112 side for the same mass of injected volatile-liquid. This could be due to the relatively early injection of volatile-liquid on the female rotor side. However, the temperature on the female rotor 114 is higher than on the main rotor 112 close to the working-fluid outlet 108 which is critical as this temperature is much higher than the temperature in the working-fluid inlet 106. Volatile-liquid Iso-surface is presented in the region where working-fluid temperature is below the saturation temperature at 11.0 bar. Specifically, if the temperature is above this limit, volatile-liquid is converted into vapor and as such is not taken in consideration further on. The experiments show that this effect is visible in the compression-chamber 104 opened to the working-fluid outlet 108, also in the working-fluid outlet 108 i.e. no liquid volatile-liquid is present here.

[00048] The experiments for calculating the vapor formation and cooling of working-fluid is conducted. The result shows the working-fluid temperature distribution on the main rotor surface (case 2 of the experimental data to observe operating parameters of the compressor as shown in FIG. 7), in the end leakage and in a plane through the working- fluid outlet 108. Also, the region where liquid volatile-liquid is getting converted to vapor. The result further shows distribution of liquid volatile-liquid on the main rotor surface and also the heat energy being removed from working-fluid in regions where evaporation is active. The working-fluid temperature and presence of liquid volatile-liquid can be correlated to the regions of vapor formation and heat extraction. Due to very low mass of volatile-liquid 0.009 Kg/sec in Case 2 the local working- fluid temperature reaches to about 290°C. The results further show evaporation in Case 3 which has 5 times higher mass injection as compared to Case 2. In this case the peak working-fluid temperature has dropped to below 200°C.

[00049] The experiments for calculating the thermal deformation of the rotors 112, 114 of the single stage rotary screw compressor 100 of FIG.l is conducted. The result of the experiments is derived by performing analysis of thermal deformation as a function of various operating conditions. The results depicted that a minimum deformation of 26.56 microns is recorded on the suction side of the female rotor 114 and the maximum deformation on the discharge side is 107.27 microns. Moreover, on the tip of the main rotor 112 the minimum deformation is 31.3 microns on the suction side and maximum is 86.4 microns on the discharge side.

[00050] Additionally, at the discharge end gap, the maximum deformation is 54.6 and 113.68 microns on the main and female rotor 112, 114 respectively. Furthermore, a maximum deformation in the order of 107.76 and 109.21 microns on the main rotor 112 and female rotor 114 respectively is observed.

[00051] The experiments for calculating the temperature distribution and resultant thermal deformation of the main rotor of the single stage rotary screw compressor 100 of FIG.l is conducted. The result of the experiments shows the temperature distribution on the main rotor 112 and the resultant thermal deformation of the main rotor lobe profile at the discharge end. Furthermore, the resultant thermal deformation of the main 112 rotor is evaluated on the female rotor lobe profile at the discharge end, this represents the interference of the rotor profiles during operation. Additionally, the magnitude of deformation indicates that the rotor profile will have considerable metal to metal interference resulting into surface damage. [00052] The experiments for calculating the temperature distribution in the two rotors of the sing le stage rotary screw compressor 100 of FIG. l is conducted . The result depicts the temperature distribution in the two rotors under a linearized boundary condition. The result further shows that the temperature distribution in the two rotors may be within a range of 50°C and 170°C.

[00053] The experiments for calculating the deformation of the rotors of the single stage rotary screw compressor 100 of FIG. l is conducted . The results of deformation calculations are obtained from a linearized temperature distribution. The results show that a minimum deformation of 33.8 microns and a maximum is 137.8 micron is obtained on the tip of the female rotor 114. Additionally, a minimum deformation is 31.49 and maximum is 106.02 microns is obtained on the tip of the main rotor 112. Moreover, maximum deformation is 89.12 and 124.63 microns on the main and female rotor 112, 114 respectively is obtained on the discharge end gap.

[00054] Furthermore, these deformation predictions indicate that conditions are not improved even with linearization of temperature distribution. Additionally, a considerably higher cooling effect would be required to be introduced into the compression process to scale down the magnitude of temperature rise in the compressor.

[00055] Optionally, the present disclosure relates to a method for determining the specific quantity of the volatile-liquid for the single stage rotary screw compressor, such as the single stage rotary screw compressor. Specifically, the specific quantity of volatile-liquid is injected in the compression-chamber during compression of the working-fluid for reducing a temperature of the compressed working-fluid to be released through the working-fluid outlet. More specifically, the specific quantity of volatile-liquid evaporates in the compression-chamber to reduce the temperature of compressed working-fluid present therein. It is to be understood that the volatile-liquid derives latent heat of evaporation from the compressed working-fluid in the compression-chamber. Thus, as the volatile-liquid injected in the compression-chamber evaporates, the temperature of the compressed working-fluid in the compression- chamber is reduced (under constant pressure conditions). Furthermore, reduction in temperature of compressed working-fluid in the compression-chamber is directly proportional to the specific quantity of volatile-liquid injected in the compression-chamber. However, it will be appreciated that the compressed working-fluid only allows evaporation of volatile-liquid till saturation of the compressed working-fluid is achieved. Consequently, the temperature of the compressed working-fluid can only be reduced (by evaporation) till the instance, saturation of the compressed working-fluid is achieved. Therefore, highest quantity of volatile-liquid that can be injected in the compression-chamber is the quantity of volatile-liquid that can be completely evaporated, thereby saturating the compressed working-fluid. In other words, compression- chamber can be injected with volatile-liquid until the compressed working-fluid is saturated and further evaporation stops. Therefore, the method relates to determining a specific quantity of volatile-liquid, to be injected in the compression-chamber, that saturates the compressed working-fluid (namely, the highest quantity of volatile-liquid).

[00056] Optionally, the present disclosure relates to determining the highest quantity of volatile-liquid to be injected in the compression- chamber to reduce the temperature of the compressed working-fluid in the compression-chamber until saturation. It is to be understood that the specific quantity of volatile-liquid injected in the compression- chamber is lesser than or equal to the highest quantity of volatile-liquid. The method for determining the highest quantity of volatile-liquid to be injected in the compression-chamber comprises an iterative computation using known input parameters to achieve working-fluid (to be released at the working-fluid outlet) having a relative humidity of 100 percent. Specifically, such iterative computation determines the highest quantity of volatile-liquid (in terms of mass), absolute humidity of working-fluid at working-fluid inlet (for saturation) and temperature of working-fluid at working-fluid inlet. Furthermore, the known input parameters include compression power, working-fluid mass flow rate, suction relative humidity, suction pressure, suction temperature and pressure at working-fluid outlet. The iterative computation comprises minimizing a value of absolute humidity of compressed working-fluid at working-fluid outlet. Specifically, an initial value of absolute humidity of compressed working-fluid at working-fluid outlet may be assumed and a final value of absolute humidity of compressed working-fluid at working-fluid outlet may be calculated. Consequently, a difference of the final value and the initial value may be minimized. Furthermore, temperature of compressed working-fluid at working-fluid outlet depends on saturation temperature at delivery pressure. Additionally, Steltz and Silvestri (1958) correlation is used to evaluate pressure of saturated steam as a function of temperature. It will be appreciated that results from the iterative computation indicate that effectiveness of working-fluid compression with volatile-liquid evaporation to cool the compressed working-fluid reduces as the suction temperature increases.

[00057] Optionally, the optimal operation of the single stage rotary screw compressor, such as the single stage rotary screw compressor, requires various parameters to be analyzed and to be calculated such as delivery temperature, delivery absolute humidity, saturation pressure, thermal deformations and so forth. Several analysis approaches utilizing various solvers are available that can be used to analyze the various parameters. For example, the various solvers that are used are ANSYS CFX solver, ANSYS FEA solver, ANSYS ICEM solver and so forth. The entire working domain of the single stage rotary screw compressors can be disintegrated into four main sub-domains such as a rotor domain, an working-fluid inlet (or suction port), a discharge end leakage gap and an working-fluid outlet (or discharge port). A 3D computational fluid dynamics (CFD) model is used to evaluate temperature distribution inside the compression-chamber. The results obtained from the 3D CFD are used as an input to analyze thermal deformation. The 3D CFD model requires a specific type of solver for analysis which affects the output of the analysis performed by the 3D CFD model. In an embodiment, the ANSYS CFX solver with a Eulerian-Eulerian approach for 3D CFD modelling was used to calculate working-fluid and volatile-liquid distribution. Furthermore, SCORG grid generator has been used to calculate the specific grid requirements for the calculation of deforming the rotor domains.

[00058] Optionally, the main parameters for the designing of screw rotor parameters that has been taken under consideration comprises an axis distance, a pitch diameter, an outer diameter, an inner diameter, a rotor length, a lead, a wrap angle, a lead angle and a helix angle. All the above-mentioned parameters have been taken into account for both the main rotor and the female rotor. Some essential leakage gaps are present such as a radial and interlobe leakage path, an axial end leakage gap and so forth. The leakage gaps have been set at 30 microns. The grid for the rotor domain has been developed by using the SCORG grid generator while the grids for all stationary domains have been generated by using the ANSYS ICEM. The working domain of the single stage rotary screw compressor has been sub divided into four major sections i.e. a suction port, a rotor domain, a discharge end leakage gap and an working-fluid outlet. The suction port of the working domain of the single stage rotary screw compressor provides the suction pressure boundary condition and the connecting interfaces with the other rotor domains. Additionally, the suction port comprises two volatile-liquid injection ports that are located on the side of the rotors of the rotor domain. Furthermore, each of the two volatile-liquid injection ports have one of their ends to be in a fixed state and another end to be in a free state. The free end of each of the volatile-liquid injection ports is connected to a source of volatile-liquid wherefrom volatile-liquid can flow into the corresponding volatile-liquid injection port. Consequently, the free end of the injection ports has been set to define the volatile-liquid mass flow rates whereas the other end of the volatile-liquid injection port is made to interface with the rotor domain. The rotor domain of the working domain of the single stage rotary screw compressor is the main element of the main compressor model. The rotor domain interfaces with the suction port, the two volatile-liquid injection tubes and the discharge end leakage gap. The numerical grid for the rotor domain has been generated by using the SCORG grid generator as stated above.

[00059] Optionally, a single block rotor domain for the multiphase flow simulations has been generated that houses the main rotor and the female rotors. Additionally, the interlobe leakage gaps and the radial leakage gaps have been included in the rotor domain meshes for analytical purposes. The rotor domain may deform with time. The discharge end leakage gap of the working domain of the single stage rotary screw compressor depicts the discharge leakage gap. The rotor and the working-fluid outlet connect to the discharge end leakage gap through a non-conformal interface. The non-conformal interface is such an interface that does not have equal no of nodes on both sides of the interfacing surfaces. The working-fluid outlet of the working domain of the single stage rotary screw compressor depicts an working-fluid outlet. The high-pressure boundary conditions have been applied on the discharge outlet (such as the working-fluid outlet). Furthermore, the working-fluid outlet has been connected to the end leakage through a non-conformal interface. The single stage rotary screw compressor model comprises a few main non-conformal surfaces such as a rotor- axial suction port, a rotor-radial suction port + volatile-liquid injection tubes, a rotor-axial discharge end leakage gap, a discharge end leakage gap-working-fluid-outlet, an working-fluid-outlet-discharge pipe and so forth. The abovementioned non-conformal interfaces have been defined with the use of a general grid interface (GGI) option and with the use of conservative flux conditions in the 3D CFD model.

[00060] Furthermore, in the 3D CFD model of the single stage rotary screw compressor, a non-overlap conditions on the rotor side of the rotor domain have been set as a No-Slip fixed wall with respect to a boundary frame. Moreover, for the 3D CDF analysis, the approach used by the solvers are mainly a Eulerian-Eulerian approach or a Eulerian-Lagrangian approach. The Eulerian-Eulerian approach is appropriate for a problem consisting high volume fraction. Furthermore, the Eulerian-Eulerian approach there considers a relative slip between the phases that can be an inhomogeneous phase or a homogeneous phase. The consideration of the inhomogeneous phase qualifies as an appropriate model for the volatile-liquid injected single stage rotary screw compressors because the inhomogeneous phase is able to treat a momentum transport for each phase separately and further the inhomogeneous phase is able to account for a condition requiring high slip for the optimal operation of the single stage rotary screw compressor. Moreover, the parameters such as an interphase heat, a mass and a momentum transfer are also required to be modelled.

[00061] Optionally, the abovementioned transfer phenomena have been entirely dependent on the characteristics of the geometry and spread of the phases within each other. Therefore, an assumption needs to be made so that the model that has been selected is applicable to the flow regime of the single stage rotary screw compressor and characteristics of the phase interaction inside the single stage rotary screw compressor. The calculation of the volatile-liquid evaporation requires a few steps. It is further required to use two sources for the multiphase formulation so as to account for the volatile-liquid evaporation. The two sources that are required are a latent heat energy source and a water vapor mass source. In one embodiment, the latent heat received from the latent heat energy source at a pressure of 11.0 bar gets activated when a local gas temperature exceeds the saturation temperature of 184.06°C. Furthermore, this latent heat energy has been eliminated from the gas phase. Additionally, the volatile-liquid mass received from the water vapor mass source gets activated and eliminated from the volatile-liquid phase.

[00062] The formulae for the calculation of the water vapor energy and the water vapor mass has been mentioned below:

Volatile-LiquidVapourEnergy = -(step (Air Ideal Gas.Temperature/l[K]- 457.21)) *1998.55[m^2sec^-2]*1000[] *Volatile-

LiquidSpray.Density*Volatile-LiquidSpray.Volume Fraction/timestep

Volatile-LiquidVapourMass = -(step (Air Ideal Gas.Temperature/l[K]- 457.21)) * Volatile-LiquidSpray. Density* Volatile-LiquidSpray. Volume

Fraction/timestep

[00063] As there is a requirement of stable operation of the single stage rotary screw compressor, there is a need to provide better initial conditions to the various problems faced in the operation of the single stage rotary screw compressor. Therefore, the discharge pressure is slowly increased (or ramped) with time. The function used for defining the linear variation of pressure with time is mentioned below:

[00064] Pressure ramp function at Outlet

PressOut = (0.0[bar]) + (step (300[]-Accumulated Time Step) *((10.0[bar]/300[]) * (Accumulated Time Step)) + step (Accumulated Time

Step-300[]) * (10.0[bar])) The number of time steps used depends on the magnitude of the pressure, the volatile-liquid mass flow and the number of compression cycles covered in the time steps.

[00065] Optionally, for obtaining the 3D CDF model of the Rotor Grid Generation SCORG rotor grid generation techniques have been taken into account. There are the set of parameters that needs to be used for the multiphase flow calculations in the ANSYS CFX solver. The multiphase flow calculations comprise certain steps such as:

[00066] Step 1 : Import rotor profiles and position them in the correct orientation such that from the engagement point to the disengagement point for one interlobe rotation, there is no intersection of the profiles. Set the geometrical inputs for clearance gaps, Relative length (L/D ratio) and Wrap angle. Set the parameter 'Number of Profile Points' to about 1200. Generate Numerical Rack. [00067] Step 2: Set Circumferential, Radial, Angular and Interlobe divisions. The number of divisions produces a reasonable node count in the rotor mesh and a reasonable time step size with the main rotor rotation of 2.4 degrees per time step.

[00068] Step 3: A method of distribution 'Rack to Rotor Conformal' is suggested for multiphase flows and has been used in all test cases. This method produces a single domain mesh for both rotors. Generate and Inspect distribution so that no irregularities are visually seen. No warnings should be reported by SCORG.

[00069] Step 4: Generate the mesh in each cross section. No errors should be reported by SCORG. Meshing parameters are applied in order to produce smooth and orthogonalized mesh in the cross section.

[00070] Step 5: Select ANSYS CFX as the pre-processor and generate 3D mesh files. [00071] The thermal deformation analysis for the single stage rotary screw compressor has been performed using the ANSYS FEA (finite element analysis) multiphase solver. Furthermore, the input parameters that are required for the thermal deformation analysis are the distribution of temperature on the boundaries of the single stage rotary screw compressor components. The thermal deformation model comprises mainly of three modules, i.e. a transfer of the boundary temperatures from the 3D CFD model to the thermal deformation model, a solving thermal conduction equation by usage of the ANSYS FEA in the solid components and a solving deformation equation by using the ANSYS FEA in the solid components. Additionally, contact pairs are established in the gap regions of the single stage rotary screw compressor in the thermal deformation model that transfer the forces and deformation from one component of the single stage rotary screw compressor to the other during analysis. Moreover, a gap size of 30 microns exists in the CAD assembly between the two rotors, between the rotors and the housing bore and between the discharge end face of the rotors and the end plate.

[00072] Optionally, a 3D computational fluid dynamics (CFD) model is used to evaluate temperature distribution inside the compression- chamber. The results obtained from the 3D CFD are used as an input to analyze thermal deformation. The present disclosure describes a general multiphase system consists of interacting phases dispersed randomly in space and time. Moreover, a Eulerian-Eulerian multiphase approach is used to model droplets or bubbles of secondary phase(s) dispersed in continuous fluid phase (primary phase). Additionally, the approach allows for mixing and separation of phases, solves momentum, enthalpy, and continuity equations for each phase and tracks volume fractions, uses a single pressure field for all phases, uses interphase drag coefficient, allows for virtual mass effect and lift forces, multiple species and homogeneous reactions in each phase, allows for heat and mass transfer between phases and can solve turbulence equations for each phase. [00073] The Eulerian multiphase equation for continuity, momentum and energy are based on the following fundamental physical principles of fluid dynamics:

(a) mass is conserved;

(b) F = ma (Newton's second law); and

(c) energy is conserved.

The Eulerian multiphase model equations are for the single stage rotary screw compressor of the present disclosure are calculated to be as follows:

(1) Continuity:

wherein a _q is volume fraction for q ^th phase.

(2) Momentum for q ^th phase

[00074] The equation for momentum includes components for transient, convection, pressure, body, shear, interphases force exchange, interphase mass exchange, external, virtual and lift mass forces.

The interphase exchange forces are expressed as:

(3) Energy equation for the q ^th can be similarly formulated .

[00075] The present disclosure relates to a simplified enthalpy balance approach. The simplified enthalpy balance approach describes working-fluid as the primary phase of interest with a very small quantity of volatile-liquid injected, volatile-liquid is the secondary phase considered for effect on heat transfer and sealing in leakage gaps. Furthermore, working-fluid and volatile-liquid multiphase flow is solved using Eulerian-Eulerian approach. Moreover, volatile-liquid evaporation is accounted by enthalpy balance. It will be appreciated that the various parameters associated with the simulation of the single stage rotary screw compressor, such as single stage rotary screw compressor, are considered and modified (for example, to include operation of the single stage rotary screw compressor under adverse conditions) while still maintaining satisfactory temperature of the compressed working-fluid and being decodable in real-time.

[00076] The results obtained by the simulations show that the single stage rotary screw compressors provide better outlet pressure, thermal efficiency, energy consumption, uniform thermal deformation of main rotor, female rotor and the compression-chamber, providing working- fluid pressure ranging up to 15 bar using single stage compression and so forth with half of the weight and complexity.

[00077] Notably, the simulations were carried out to study the feasibility of the single stage rotary screw compressor to for releasing the working-fluid though the working-fluid outlet at 11.0 bar pressure with very low quantity of volatile-liquid injection in a manner that the quantity of volatile-liquid substantially cools compressed working-fluid .

[00078] In an embodiment, the pair of rotors rotate at a speed in a range of 1000 rotations per minute (rpm) to 4000 rpm. The pair of rotors can be configured for operation at low rotation speeds, such as, rotation speeds of less than 4000 rpm. Results obtained using the aforementioned parameter (i.e. provid ing 11.0 bar pressure) shows that hig her cooling is obtained when the main rotor and female rotor are rotating at a speed of 4500 rpm than at a speed of 6000 rpm. In such a simulation, it is to be understood that for the same mass flow rate, total mass of volatile-liquid injected and residence time of the injected volatile- liquid is higher at lower speed. Since the residence time of injected volatile-liquid is higher at lower speed, greater heat transfer between the compressed working-fluid and the injected volatile-liquid can be achieved. Furthermore, at 4500 rpm the working-fluid mass flow rate is lower than at 6000 rpm which result in different power requirement. Therefore, the same mass of volatile-liquid will provide higher cooling at lower speeds. For example, such lower speeds can include rotation speeds in the range of 1000 rpm to 4000 rpm. Furthermore, as shown in case 1 FIG. 7, when the mass of volatile-liquid required is doubled than the mass required for saturation is injected in the compression-chamber, the discharge temperature exceeds 300°C. Flowever, for a low volatile- liquid mass flow rate of 0.009 Kg/sec the cooling effect is stronger in Case 2 as compared to Case 1 which has 2x volatile-liquid mass flow. This is due to a combined effect of increase in working-fluid flow and decrease in residence time at 6000rpm. Moreover, as shown in case 3, injecting 5 times higher volatile-liquid mass flow at 4500 rpm, the temperature of approximately 200°C is achieved. Furthermore, it was found during the simulations that cooling effect of the injected volatile-liquid is higher on the female rotor side due to early injection as well as lower compression temperature rise thereon. Flowever, the peak temperature close to the working-fluid outlet is higher on the female rotor. Additionally, when the temperature of the main rotor and female rotor is below 200°C, the average deformation exceeds 40 microns on a root of the main rotor and female rotor, and 90 microns on the tips of the main rotor and female rotor. In such a case, deformation exceeds 90 microns near the working- fluid outlet. Such deformations can be overcome by using the aforementioned cooling jacket (mentioned in FIG. 1).

[00079] Optionally, based on the simulations, it was found that when an equal mass of volatile-liquid is injected at the main and female rotor side, the cooling effect is greater on the female rotor. An increase in the volatile-liquid injection on main rotor can facilitate better temperature uniformity.

[00080] Optionally, the volatile-liquid injecting arrangement is arranged on the in proximity of the working-fluid inlet. In such a case, when the temperature of the injected volatile-liquid is 10°C, the volatile- liquid injecting arrangement cools the working-fluid entering the compressor chamber and, also provide increased rotor film formation that can help cooling and lubrication of the rotors.

[00081] Optionally, the main rotor, the female rotor and the compression-chamber can be manufactured using a material having a low thermal expansion coefficient. More optionally, the main rotor, the female rotor and the compression-chamber is coated with a material having the low thermal expansion coefficient. In an example, the housing and the pair of rotors are made of a material selected from a group consisting of Invar, cast iron or any combination thereof. More optionally, the material is heat-treated and/or coated using a different material. For example, the housing and the pair of rotors are made from heat-treated steel. In another example, the housing and the pair of rotors are made from steel and are coated with zinc. Alternatively, examples of such materials, may include but are not limited to molybdenum, tungsten, silicon carbide and silicon nitride.

[00082] Furthermore, the single stage rotary screw compressor can operate at temperature ranging up to 500°C, while maintaining minimal clearance, without any seizures. The main rotor and the female rotor drive each other directly without any use of synchronizing gears.

Moreover, adiabatic efficiency of the single stage rotary screw compressor is achieved using reduced leakage, reduced thermal expansion and providing a mechanical sealing between the compressor chamber and a bearing housing. Such an arrangement prevents the ingress of the grease into the compressor chamber and further reduces the leakage between the main rotor, the female rotor and compressor chamber.

[00083] The present disclosure also relates to the method as described above. Various embodiments and variants disclosed above apply mutatis mutandis to the method.

[00084] A method of manufacturing a single stage rotary screw compressor, the method comprising providing a housing having an working-fluid inlet, a compression-chamber, and an working-fluid outlet, wherein the working-fluid inlet and the working-fluid outlet are fluidically coupled to the compression-chamber; arranging a rotary-screw mechanism within the compression-chamber, the rotary-screw mechanism comprising a pair of rotors defining an inlet-end and an outlet-end of the rotary-screw mechanism, wherein the rotary-screw mechanism draws working-fluid into the compression-chamber through the working-fluid inlet, receives the working-fluid between the pair of rotors from the inlet-end, compresses the working-fluid between the pair of rotors, releases the compressed working-fluid from the outlet-end and discharges the compressed working-fluid from the compression-chamber from the working-fluid outlet; providing a volatile-liquid injecting arrangement having at least one volatile-liquid injecting point arranged on the compression-chamber near the outlet-end of the rotary-screw mechanism, wherein the at least one volatile-liquid injecting point provides a specific quantity of volatile-liquid around the outlet-end of the rotary-screw mechanism to allow the specific quantity of volatile-liquid to evaporate and reduce a temperature of the compressed working-fluid released through the outlet-end; and evaporate and reduce a temperature of the pair of rotors at the outlet-end, to reduce a thermal deformation experienced by each rotor of the pair of rotors during compression of the working-fluid therebetween; and arranging a cooling jacket on the housing near the outlet-end of the rotary-screw mechanism, wherein the cooling jacket reduces the temperature of the pair of rotors at the outlet-end, to reduce the thermal deformation experienced by each rotor of the pair of rotors.

[00085] Optionally, the method further comprises arranging a bearing arrangement within the housing, wherein the bearing arrangement comprises a pair of first-bearings to rotatably support the inlet-end of each rotor of the pair of rotors and a pair of second-bearings to rotatably support the outlet-end of each rotor of the pair of rotors.

[00086] Optionally, the method comprises arranging the cooling jacket on the housing to enclose the pair of second-bearings.

[00087] Optionally, the at least one volatile-liquid injecting point provides the specific quantity of volatile-liquid to reduce a temperature of the compression-chamber.

[00088] Optionally, the specific quantity of volatile-liquid is less than or equal to a quantity of volatile-liquid required for saturation of the compressed working-fluid at the working-fluid outlet.

[00089] Optionally, the method further comprises providing a non return valve fluidically coupled with the working-fluid outlet, for reducing the temperature of the compressed working-fluid at the working-fluid outlet by regulating a pressure thereof.

[00090] Optionally, the method further comprises providing an working-fluid dryer fluidically coupled with the non-return valve, wherein the working-fluid dryer removes evaporated volatile-liquid from the compressed working-fluid discharged through the working-fluid outlet via the non-return valve.

[00091] Optionally, the housing and the pair of rotors are manufactured using : Invar, steel, cast iron, iron alloy, aluminum alloy or a combination thereof and the material is heat-treated and/or coated using a different material..

[00092] Optionally, a separation between the rotors of the pair of rotors is in a range of 10 microns to 50 microns; and a separation between the pair of rotors and the compression-chamber is in a range of 10 microns to 50 microns.

[00093] Optionally, the pair of rotors rotate at a speed in a range of 1000 rotations per minute (rpm) to 4000 rpm.

[00094] The present disclosure provides the single stage rotary screw compressor 100 and the method of manufacturing thereof as explained herein above. The single stage rotary screw compressor 100 comprises the volatile-liquid injecting arrangement having at least one volatile- liquid injecting point 140 and 142. Such volatile-liquid injecting arrangement having at least one volatile-liquid injecting point provides the specific quantity of volatile-liquid near the outlet-end of the rotary- screw mechanism 110 in the single stage rotary screw compressor 100 to evaporate and reduce a temperature of the compressed working-fluid released through the outlet-end 118, reduce a temperature of the pair of rotors (i.e. the main rotor 112 and the female rotor 114) at the outlet- end 118 and reduce the temperature of the compression-chamber 104. The reduction in temperature enables a reduction in thermal deformation experienced by rotating components such as the pair of rotors (i.e. the main rotor 112 and the female rotor 114). Furthermore, injection of volatile-liquid into the compression-chamber 104 enables reduction of temperature in the compression-chamber 104, thereby enabling reduction of thermal deformation occurring therein. It will be appreciated that the reduction of thermal deformation of the rotary-screw mechanism 110 (i.e. the main rotor 112 and the female rotor 114) and the compression-chamber 104 enables improvement in efficiency and increase in operating life of the single stage rotary screw compressor 100. Moreover, the cooling jacket 150 helps in reducing the temperature of the bearing arrangement (such as the pair of first-bearings 130 and the pair of second-bearings 132), thereby minimizing a requirement of reintroduction of lubricant into the bearing arrangement. Such minimization of the requirement of reintroduction of lubricant enables reduction of leakage of the lubricant into the compression-chamber 104, thus, further improving the efficiency of the single stage rotary screw compressor 100. Such single stage rotary screw compressor 100 helps in achieving temperature uniformity therein during compression of working-fluid and also reduction in the temperature of the compressed working-fluid discharged from the single stage rotary screw compressor 100. Therefore, the present disclosure provides the single stage rotary screw compressor 100 that overcomes the problems associated with high temperatures in the compressor, and enables efficient, cost effective and reliable operation of the single stage rotary screw compressor 100.

[00095] Modifications to embodiments of the present disclosure described in the foregoing are possible without departing from the scope of the present disclosure as defined by the accompanying claims. Expressions such as "including", "comprising", "incorporating", "have", "is" used to describe and claim the present disclosure are intended to be construed in a non-exclusive manner, namely allowing for items, components or elements not explicitly described also to be present. Reference to the singular is also to be construed to relate to the plural.