GB2575464A

GB2575464A - Single stage rotary screw compressor

Info

Publication number: GB2575464A
Application number: GB1811293.8A
Authority: GB
Inventors: Stupple Graham; Jacklin Julia; Jacklin Jurgen
Original assignee: Jacklin GmbH
Current assignee: Jacklin GmbH
Priority date: 2018-07-10
Filing date: 2018-07-10
Publication date: 2020-01-15
Also published as: GB201811293D0; WO2020012350A1

Abstract

A single stage rotary screw compressor 100 comprises a housing 102, with a compression chamber 104, and an air-inlet 106 and an air-outlet 108 fluidly coupled to the compression chamber. The compressor also comprises a rotary-screw mechanism 110 arranged within the compression chamber that has a pair of rotors 112, 114 each having inlet and outlet ends. A bearing arrangement 130, 132 is arranged within the housing and adjacent to the compression chamber. A water injecting arrangement has a water injecting point(s) 140, 142, arranged on a portion 144 of the compression chamber enclosing the outlet ends of the rotors. A cooling jacket 150 is arranged on a portion 152 of the housing enclosing the pair of second bearings and a non-return valve 160 is in fluid communication with the air outlet for regulating a pressure within the compression chamber during compression of the air. The compressor may comprise an air dryer in fluid communication with the non-return valve to move evaporated water content from the compressed air released through the air outlet. Preferably the housing and rotors are made from material which may include invar, cast iron or a combination thereof.

Description

SINGLE STAGE ROTARY SCREW COMPRESSOR

TECHNICAL FIELD

The present disclosure relates generally to air-compressors; and more specifically, to a single stage rotary screw compressor operable to provide a temperature controlled compressed air.

BACKGROUND

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.

However, there are various limitations associated with the conventional air compressors (for example, with rotary-screw compressors), for example primarily related to managing a temperature within the compressor and to control a temperature of the compressed air to be generated by the compressor. Furthermore, such management of temperature may be read in line or associated with other problems, such as, high internal leakage, nonuniform thermal deformation of rotors and the compression chamber, only a limited delivery pressure can be achieved using the single stage compression and so forth. Furthermore, if the single stage compression is used to generate high pressurized air, it has its own disadvantages such as excessively high operational temperature, severe mechanical and thermal stress subjected to various compressor components and so forth. Moreover, the multistage compression requires an intercooling arrangement, which adds to an operational cost thereof.

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

SUMMARY

The present disclosure seeks to provide a single stage rotary screw compressor operable to provide a temperature controlled compressed air.

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

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

- a rotary-screw mechanism arranged within the compression-chamber, the rotary-screw mechanism comprises a pair of rotors, each having an inlet-end and an outlet-end, wherein the rotary-screw mechanism is operable to draw air into the compression-chamber through the air-inlet, compress the drawn air within the compression-chamber and release the compressed air through the air-outlet;

- a bearing arrangement arranged within the housing and adjacent to the compression-chamber, the bearing arrangement comprises a pair of firstbearings operable to rotatably support the inlet-end of each of the pair of rotors and a pair of second-bearings operable to rotatably support the outletend of each of the pair of rotors;

- a water injecting arrangement having at least one water injecting point arranged on a portion of the compression-chamber enclosing the outlet-ends of the pair of rotors, wherein the at least one water injecting point is operable to provide a specific quantity of water around the outlet-end of the pair of rotors to allow the specific quantity of the water to evaporate during compression of the air for reducing a temperature of the compressed air released through the air-outlet;

- a cooling jacket arranged on a portion of the housing enclosing the pair of second-bearing, the cooling jacket is operable to reduce temperature around the outlet-ends of the pair of rotors; and

- a non-return valve in fluid communication with the air-outlet for regulating a pressure within the compression-chamber during compression of the air.

The present disclosure provides an at least partial solution to the aforementioned technical problem, or problems, associated with known art, wherein the water injecting arrangement enables the single stage rotary screw compressor to generate a temperature controlled compressed air. Further, the single stage rotary screw compressor of the present disclosure precludes a use of oil and a need for multiple stages to generate the temperature controlled compressed air having a desired pressure. Moreover, the cooling jacket enables in manging thermal deformation aspects of various components of the compressor. Additionally, the non-return valve also enables in controlling temperature of the compressed air to be generated by the compressor.

In another aspect, an embodiment of the present disclosure provides a method for determining the specific quantity of the water for the aforementioned single stage rotary screw compressor.

The present disclosure provides a single stage rotary screw compressor with a simplified and economic design and provides a desired pressurised output of air with reduced temperature. Notably, the desired pressurised output of the air is provided without intervention of multiple stages. Specifically, the single stage rotary screw compressor involves a single stage process and does not employ multiple stages for compression and cooling of the air. Beneficially, the single stage rotary screw compressor comprises a light-weight assembly. Furthermore, the thermal efficiency of the single stage rotary screw compressor is substantially high.

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.

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

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. However, 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.

Embodiments of the present disclosure will now be described, by way of example only, with reference to the following diagrams wherein:

FIG. 1 is a block diagram 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 disclosure;

FIG. 5 is a schematic illustration of a rack curve configuration between the rotors, in accordance with an embodiment of the present disclosure;

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 water 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 water mass requirement on varying the compression power of the single stage rotary screw compressor, in accordance with an embodiment of the present disclosure;

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;

FIG. 17 are exemplary visual representations of experimented cases for air temperature distribution (relates to the experimental data to observe operating parameters of the air compressor as shown in FIG. 7), in accordance with an embodiment of the present disclosure;

FIG. 18 are exemplary visual representations of experimented cases for air temperature distribution (relates to the experimental data to observe operating parameters of the air compressor as shown in FIG. 7), in accordance with an embodiment of the present disclosure;

FIGs. 19 and 20 are exemplary visual representations of experimented cases for vapour formation and cooling of air (relates to the experimental data to observe operating parameters of the air compressor as shown in FIG. 7), in accordance with an embodiment of the present disclosure;

FIGs. 21 and 22 are exemplary visual representations of thermal deformation of the rotors of the single stage rotary screw compressor of FIG.l, in accordance with an embodiment of the present disclosure;

FIG. 23 is an exemplary visual representation of temperature distribution and resultant thermal deformation of the main rotor of the single stage rotary screw compressor of FIG.l, in accordance with an embodiment of the present disclosure;

FIG. 24 is an exemplary visual representation of temperature distribution in the two rotors of the single stage rotary screw compressor of FIG.l, in accordance with an embodiment of the present disclosure; and

FIG. 25 is an exemplary visual representation of results of deformation calculations of the rotors of the single stage rotary screw compressor of FIG.l, in accordance with an embodiment of the present disclosure.

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

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 practising the present disclosure are also possible.

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

Optionally, the single stage rotary screw compressor further comprises an airdryer operable to remove evaporated water content from the compressed air released through the air-outlet.

Optionally, 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

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 having an air-inlet 106 and an air-outlet 108. The air-inlet 106 and the air-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 rotaryscrew 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 inlet-end 116 in proximity to the air-inlet 106, and an outlet-end, such outlet-end 118 in proximity to the air-outlet 108.

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 air into the compression-chamber 104 through the air-inlet 106, compress the drawn air within the compression-chamber 104 and release the compressed air through the air-outlet 108.

The single stage rotary screw compressor 102 further comprises a bearing arrangement arranged within the housing 102 and adjacent to the compression-chamber 104. 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 firstbearings 130 operable 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 operable 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.

The single stage rotary screw compressor 100 further comprises a water injecting arrangement having at least one water injecting point arranged on a portion of the compression-chamber enclosing the outlet-ends of the pair of rotors. As shown, the water injecting arrangement includes the at least one water injecting point, such as a pair of holes 140, 142 (or channel) a portion 144 of the compression-chamber 104 enclosing the outlet-ends 118 of the pair of rotors (i.e. the main rotor 112 and the female rotor 114).

The at least one water injecting point, such as pair of holes 140, 142, is operable to provide a specific quantity of water around the outlet-end 118 of the pair of rotors (i.e. the main rotor 112 and the female rotor 114) to allow the specific quantity of the water to evaporate during compression of the air for reducing a temperature of the compressed air released through the airoutlet 108.

It may be appreciated that the water injecting arrangement may further include a water source, pipes (or conduits) coupled to the at least one water injecting point (such as a pair of holes 140, 142) and a pump (not shown), which may be instructed for pumping or injecting the specific quantity of water into the compression-chamber 104 through the pair of holes 140, 142.

In one embodiment, the single stage rotary screw compressor 100 further a cooling jacket 150 arranged on a portion 152 of the housing 102 enclosing the pair of second-bearing 132. The cooling jacket 150 acts as a heatexchanger for reducing the temperature of various components of the single stage rotary screw compressor 100, thereby allowing reducing the deformation occurring therein. In the present embodiment, the cooling jacket 150 is operable to reduce temperature around the outlet-ends 118 of the pair of rotors (i.e. the main rotor 112 and the female rotor 114. In an example, the cooling jacket 150 may be hollow housing, arranged on the portion 152 of the housing 102, that receives a coolant, such as water 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 air-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.

According to an embodiment, the single stage rotary screw compressor 100 further comprises a non-return valve 160 in fluid communication with the airoutlet 108 for regulating a pressure within the compression-chamber 104 during compression of the air. The non-return valve 160 facilitates in reduction of temperature for the compressed air exiting the air outlet 106.

For example, the non-return valve 160 prevents pressure from returning into the compression chamber 104 and exhausting through the air inlet 104 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 air exiting the air-outlet 108 for a certain time incessant. This results in a degree of pressure relief in the air-outlet 108 and consequently a reduced temperature thereof.

In one embodiment, the single stage rotary screw compressor 100 further comprises an air-dryer 170 in fluid communication with the non-return valve 160. The air-dryer is operable to remove evaporated water content from the compressed air released through the air-outlet 108. It will be appreciated that the process of air compression concentrates atmospheric contaminants, including water vapour. This raises the dew point of the compressed air relative to free atmospheric air and leads to condensation within the air-outlet 108 and the compressed air cools downstream of the compressor 100.

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 air-inlet 106, the air-outlet 108, the main rotor 112, the female rotor 114, and the at least one water injecting point (such as the pair of holes 140, 142).

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. According to an embodiment, the main rotor 112 and the female rotor 114, i.e. outer surfaces thereof are spaced apart by a distance of 30 microns. Also, 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).

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 Ti,₇(x,y). Furthermore, points on the rotor profile (i.e. female rotor 114) are r_iJ=o(x,y) and points on outer boundary consisting of casing and rack curve are r_iJ=i(x,y) and the point distribution on outer full circle is r_i;-,₌₁(x,y). Each background blocks is identified by its index £?,. Further, the points on the inner boundary of the blocks which are the rotor profile nodes are rb_ij=0(x,y) and point distribution on outer full circle is rbij'^x.y).

The nodes are distributed on the outer circle covering the rack part with required number of points i_rack. Further, the nodes are distributed on the outer circle covering the casing part with required number of points i_caSi_ng· Subsequently, the nodes the data are available for rb_iJ=0(x,y), rb_ij>₌₁(x,y') and r_i;/₌₁(x,y) and is required to calculate η,₇₌₀(^τ)· The node distribution is based on equidistant spacing as given in equation below:

n,r=i(^x>y) = u-i,₇'=i(uy)+ wherein, = — and

Si= r_iJ>₌₁(.x,y')- r_i=oj'₌₁(x,y)

A scanning function has the information of the background blocking. The scanning function traces each node r_i;/₌₁(x,y) and identifies the block B_t 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 η,₇=₀(χ,γ) 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_t. Further, as that of the outer circle nodes r_u/₌₁(x,y). Further, the illustration provides projection of r_i;/₌₁(x,y) on the inner boundary of the block to get η,₇=₀(χ,γ), wherein the projection is based on arc length factor given by equation below:

Si ri,j=o(x,y) = rb_iJ=0(x,y) + (rh_i+lj7=0(x,y) - rb_u=0(x,y)) — wherein, S, = r_i7-/₌₁(x,y) — rb_ij/₌₁(x,y') ^si = rb_i+1J:₌₁(x,y) - r\₇'₌₁(x,y)

Further, the calculated positions ri,₇=₀(x,y) of nodes ensure that they are guided by a regular rotor profile. Further, regularised distribution is superimposed onto the rack curve by finding the intersection points of the distribution lines and the rack curve.

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 r_i7=1(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.

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. However, 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 air, 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 air) 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 analysed and are discussed in conjunction with FIG. 7, discussed herein later.

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 water required for saturation of the compressed air with the variation in the supplied power. In an example, the operating parameters are observed when the required pressure of the compressed air is 11 bars. As shown, the table depicts various cases, namely case 1, case 2, case 3 and case 4, at various speeds, pressures, water masses, and saturation masses and average discharged temperatures. In an example, considering low water 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 water mass flow compared to Case 2. In another example, case 3 and case 4 are designed such that the mass flow rate of water 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 water.

Referring to FIG. 8, illustrated is a graphical representation of pressure variation in the compression chamber 104 for the experimented cases of FIG. 7, 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 when the main rotor 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 water 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 air-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 350degree 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.

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 water 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.

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 1000 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 water 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 water 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.

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).

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 air 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. Furthermore, the magnitude of deformation shown in the graph 1200 indicates that the rotor profiles will have considerable metal to metal interference resulting into surface damage.

Referring to FIG. 13, illustrated is a graphical representation of experimented cases for observing water 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 water 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 water 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 water mass requirement.

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 water mass.

Referring to FIG. 15, illustrated is a graphical representation of experimented cases for observing relative water 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 water 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 water mass is required to produce saturation with 15% R.H (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.

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 water mass.

Referring to FIG. 17, illustrated are exemplary visual representations 1700 (such as a simulated heat map) of experimented cases for air temperature distribution (relates to the experimental data to observe operating parameters of the air compressor as shown in FIG. 7), in accordance with an embodiment of the present disclosure. As shown in case 2 of FIG. 7, the required water mass of 0.009 Kg/sec to achieve saturated air at the air-outlet 106 with power dissipation of 30 kW approximately. However, the visual representations 1700 herein depicts that the abovementioned estimates that do not account for transient and leakage affects. Additionally, CFD calculation has therefore resulted in higher than saturation exit temperatures.

Referring to FIG. 18, illustrated are exemplary visual representations 1800 of experimented cases for air temperature distribution (relates to the experimental data to observe operating parameters of the air compressor as shown in FIG. 7), in accordance with an embodiment of the present disclosure. As shown, the visual representations 1800 relates to experimented cases for air temperature distribution inside the single stage rotary screw compressor 100. The visual representation 1800 represents an iso-surface generated with liquid volume fraction of 0.0001. Furthermore, the temperature in the air-inlet 104 is lower on the female rotor side, but on the main rotor side shows higher air 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 water. This could be due to the relatively early injection of water on the female rotor side. However, the temperature on the female rotor 114 is higher than on the main rotor 112 close to the air-outlet 108 which is critical as this temperature is much higher than the temperature in the air inlet 106. Water Iso-surface is presented in the region where air temperature is below the saturation temperature at 11.0 bar as it is depicted in the visual representations 1800. Specifically, if the temperature is above this limit, water is converted into vapour and as such is not taken in consideration further on. This effect is visible in the compression chamber 104 opened to the air-outlet 108, also in the air-outlet 108 i.e. no liquid water is present here.

Referring to FIGs. 19 and 20, illustrated are exemplary visual representations of experimented cases for vapour formation and cooling of air (relates to the experimental data to observe operating parameters of the air compressor as shown in FIG. 7), in accordance with an embodiment of the present disclosure. As shown, visual representations 1900 shows the air temperature distribution on the main rotor surface (case 2 of the experimental data to observe operating parameters of the air compressor as shown in FIG. 7), in the end leakage and in a plane through the air-outlet 108. Also, the region where liquid water is getting converted to vapour. The distribution of liquid water on the main rotor surface and shows the heat energy being removed from air in regions where evaporation is active. The air temperature and presence of liquid water can be correlated to the regions of vapour formation and heat extraction. Due to very low mass of water 0.009 Kg/sec in Case 2 the local air temperature reaches to about 290°C. In visual representations 2000, shown the visualization of evaporation in Case 3 which has 5 times higher mass injection as compared to Case 2. In this case the peak air temperature has dropped to below 200°C.

Referring to FIGs. 21 and 22, are exemplary visual representations 2100, 2200 of thermal deformation of the rotors 112, 114 of the single stage rotary screw compressor 100 of FIG.l, in accordance with an embodiment of the present disclosure. As shown, the FIG. 21 and FIG. 22 depict the plurality of probes in located along the rotors 112, 114 to indicate the level of local deformations. Furthermore, the FIG. 21 and FIG. 22 graphically describes the results derived upon performing analysis thermal deformation as a function of various operating conditions. Specifically, the FIG. 21 shows the discharge side of the rotors 112, 114. Furthermore, 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. 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. Specifically, the FIG. 22 shows the rotors 112, 114 from the suction side. 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 shown.

Referring to FIG. 23, illustrated is an exemplary visual representation 2300 of temperature distribution and resultant thermal deformation of the main rotor of the single stage rotary screw compressor 1OO of FIG.l, in accordance with an embodiment of the present disclosure. As shown, the FIG. 23 depicts 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.

Referring to FIG. 24, illustrated is an exemplary visual representation of temperature distribution in the two rotors of the single stage rotary screw compressor 100 of FIG.l, in accordance with an embodiment of the present disclosure. As shown, the FIG. 24 depicts the temperature distribution in the two rotors under a linearized boundary condition. Furthermore, the temperature distribution in the two rotors can be within a range of 50°C and 170°C.

Referring to FIG. 25, illustrated is an exemplary visual representation of results of deformation calculations of the rotors of the single stage rotary screw compressor 100 of FIG.l, in accordance with an embodiment of the present disclosure. As shown, the FIG. 25 depicts the results of deformation calculations obtained from a linearized temperature distribution. Furthermore, 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. Optionally, a higher thermal variation on both the rotors can be seen if the results of deformation of FIG. 32 is be compared to thermal deformation as shown in FIG. 24, namely a maximum of 26.56 microns and a minimum of 107.27 microns on the tip of the female rotor 114, a maximum of 31.3 microns and a minimum of 86.4 microns on the tip of the main rotor 112, and a maximum of 54.6 and a minimum of 113.68 microns on the main and female rotor 112, 114 respectively of the discharge end gap. 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.

Optionally, the present disclosure relates to a method for determining the specific quantity of the water for the single stage rotary screw compressor, such as the single stage rotary screw compressor 100. Specifically, the specific quantity of water is injected in the compression-chamber during compression of the air for reducing a temperature of the compressed air to be released through the air-outlet. More specifically, the specific quantity of water evaporates in the compression-chamber to reduce the temperature of compressed air present therein. It is to be understood that the water derives latent heat of evaporation from the compressed air in the compressionchamber. Thus, as the water injected in the compression-chamber evaporates, the temperature of the compressed air in the compressionchamber is reduced (under constant pressure conditions). Furthermore, reduction in temperature of compressed air in the compression-chamber is directly proportional to the specific quantity of water injected in the compression-chamber. However, it will be appreciated that the compressed air only allows evaporation of water till saturation of the compressed air is achieved. Consequently, the temperature of the compressed air can only be reduced (by evaporation) till the instance, saturation of the compressed air is achieved. Therefore, highest quantity of water that can be injected in the compression-chamber is the quantity of water that can be completely evaporated, thereby saturating the compressed air. In other words, compression-chamber can be injected with water until the compressed air is saturated and further evaporation stops. Therefore, the method relates to determining a specific quantity of water, to be injected in the compressionchamber, that saturates the compressed air (namely, the highest quantity of water).

Optionally, the present disclosure relates to determining the highest quantity of water to be injected in the compression-chamber to reduce the temperature of the compressed air in the compression-chamber until saturation. It is to be understood that the specific quantity of water injected in the compressionchamber is lesser than or equal to the highest quantity of water. The method for determining the highest quantity of water to be injected in the compression-chamber comprises an iterative computation using known input parameters to achieve air (to be released at the air outlet) having a relative humidity of 100 percent. Specifically, such iterative computation determines the highest quantity of water (in terms of mass), absolute humidity of air at air inlet (for saturation) and temperature of air at air inlet. Furthermore, the known input parameters include compression power, air mass flow rate, suction relative humidity, suction pressure, suction temperature and pressure at air outlet. The iterative computation comprises minimizing a value of absolute humidity of compressed air at air outlet. Specifically, an initial value of absolute humidity of compressed air at air outlet may be assumed and a final value of absolute humidity of compressed air at air outlet may be calculated. Consequently, a difference of the final value and the initial value may be minimized. Furthermore, temperature of compressed air at air 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 air compression with water evaporation to cool the compressed air reduces as the suction temperature increases.

Optionally, the optimal operation of the single stage rotary screw compressor, such as the single stage rotary screw compressor, requires various parameters to be analysed 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 analyse 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 air-inlet (or suction port), a discharge end leakage gap and an airoutlet (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 analyse 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 air and water distribution. Furthermore, SCOR.G grid generator has been used to calculate the specific grid requirements for the calculation of deforming the rotor domains.

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 SCOR.G 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 airoutlet. 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 water injection ports that are located on the side of the rotors of the rotor domain. Furthermore, the two water injection ports have one of their ends to be in a fixed state and another end to be in a free state. Moreover, the free end of the injection ports boundaries has been set to define the water mass flow rates whereas the other end of the water 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 water injection tubes and the discharge end leakage gap. The numerical grid for the rotor domain has been generated by using the SCOR.G grid generator as stated above.

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 air-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 air-outlet of the working domain of the single stage rotary screw compressor depicts an air-outlet. The high-pressure boundary conditions have been applied on the discharge outlet. Furthermore, the air-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 + water injection tubes, a rotor - axial discharge end leakage gap, a discharge end leakage gap - air-outlet, an air-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. 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 water 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 is also required to be modelled.

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 water evaporation requires a few steps. It is further required to use two sources for the multiphase formulation so as to account for the water evaporation. The two sources that are required are a latent heat energy source and a water vapour 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 water mass received from the water vapour mass source gets activated and eliminated from the water phase.

The formulae for the calculation of the water vapour energy and the water vapour mass has been mentioned below:

WaterVapourEnergy = -(step (Air Ideal Gas.Temperature/l[K]-457.21)) * 1998.55[m^/K2sec^/K-2] *1000[] * WaterSpray. Density*WaterSpray. Volume Fraction/timestep

WaterVapourMass = -(step (Air Ideal Gas.Temperature/l[K]-457.21)) *WaterSpray.Density*WaterSpray.Volume Fraction/timestep

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:

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 water mass flow and the number of compression cycles covered in the time steps.

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:

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.

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.

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.

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 orthogonalised mesh in the cross section.

Step 5: Select ANSYS CFX as the pre-processor and generate 3D mesh files.

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.

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 analyse 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. 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:

. \ V¹ .

h 7 · [(XqPqUq) — ) . ^mpq p = l wherein a_q is volume fraction for q^th phase.

(2) Momentum for q^th phase n

— —OlqVp + CtqPqg + F · Tg + ' (Rpq T ^pq^q) + ^aqPq(Fq + Gift t,q + Fvm,q) p = l

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:

Rpq ~ Kpq(llp — Uq) (3) Energy equation for the q^th can be similarly formulated.

The present disclosure relates to a simplified enthalpy balance approach. The simplified enthalpy balance approach describes air as the primary phase of interest with a very small quantity of water injected, water is the secondary phase considered for effect on heat transfer and sealing in leakage gaps. Furthermore, air and water multiphase flow is solved using Eulerian-Eulerian approach. Moreover, water 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 air and being decodable in real-time.

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 air pressure ranging up to 15 bar using single stage compression and so forth with half of the weight and complexity.

Notably, the simulations were carried out to study the feasibility of the single stage rotary screw compressor to for releasing the air though the air-outlet at 11.0 bar pressure with very low quantity of water injection in a manner that the quantity of water substantially cools compressed air. Results obtained using the aforementioned parameter (i.e. providing 11.0 bar pressure) shows that higher 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 water injected and residence time of the injected water is higher at lower speed. Since the residence time of injected water is higher at lower speed, greater heat transfer between the compressed air and the injected water can be achieved. Furthermore, at 4500 rpm the air mass flow rate is lower than at 6000 rpm which result in different power requirement. Therefore, the same mass of water will provide higher cooling at lower speeds. Furthermore, as shown in case 1 FIG. 7, when the mass of water required is doubled than the mass required for saturation is injected in the compression chamber, the discharge temperature exceeds 300°C. However, for a low water mass flow rate of 0.009 Kg/sec the cooling effect is stronger in Case 2 as compared to Case 1 which has 2x water mass flow. This is due to a combined effect of increase in air flow and decrease in residence time at 6000rpm. Moreover, as shown in case 3, injecting 5 times higher water 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 water is higher on the female rotor side due to early injection as well as lower compression temperature rise thereon. However, the peak temperature close to the air-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 airoutlet 108. Such deformations can be overcome by using the aforementioned cooling jacket (mentioned in FIG. 1).

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

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

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. Alternatively, examples of such materials, may include but are not limited to molybdenum, tungsten, silicon carbide and silicon nitride.

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.

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.

Claims

1. A single stage rotary screw compressor comprising:

2. A single stage rotary screw compressor according to claim 1, further comprising an air-dryer in fluid communication with the non-return valve, wherein the air-dryer is operable to remove evaporated water content from the compressed air released through the air-outlet.

3. A single stage rotary screw compressor according to claims 1 or 2, wherein 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.